GIFT  OF 
Mrs.  Paul  Christiansen 


Engineering  Library 


May,  1940 


LIBRARY 

OF 

POWER  PLANT  PRACTICE 


STEAM  POWER  PLANT 
AUXILIARIES  AND  ACCESSORIES 

TERRELL  CROFT  EDITOR 


CONTRIBUTORS 

The  following    have    contributed   manuscript  or  data  or   have    otherwise 
assisted  in  the  preparation  of  this  work: 

EDMOND   SIROKT 

A.  J.  DIXON  E.  R.  POWELL 

I.  V.  LEBow  I.  O.  ROYSE 

A.  C.  STALEY — Condensers  and  Methods  of  Recooling  Condensing  Water 
JULIUS  WOLF — Injectors 


BOOKS  ON  PRACTICAL 
ELECTRICITY  , 

BY  TERRELL  CROFT 

AMERICAN  ELECTRICIANS'  HANDBOOK 

WIRING  OF  FINISHED  BUILDINGS 

WIRING  FOR  LIGHT  AND  POWER 

ELECTRICAL  MACHINERY 

PRACTICAL  ELECTRIC  ILLUMINATION 

PRACTICAL  ELECTRICITY 

CENTRAL  STATIONS 

LIGHTING  CIRCUITS  AND  SWITCHES 

ALTERNATING-CURRENT  ARMATURE  WINDING 

CONDUIT  WIRING 

ELECTRICAL       MACHINERY       AND       CONTROL 

DIAGRAMS 

CIRCUIT  TROUBLES  AND  TESTING 
ELECTRICAL-MACHINERY  ERECTION 


POWER  PLANT  SERIES 

TERRELL  CROFT 

Editor-in-chief 

STEAM  BOILERS 

STEAM-POWER-PLANT  AUXILIARIES 
STEAM-ENGINE  PRINCIPLES  AND  PRACTICE 
STEAM-TURBINE  PRINCIPLES  AND  PRACTICE 
MACHINERY  FOUNDATIONS  AND  ERECTION 
PRACTICAL  HEAT 

McGRAW-HILL  BOOK  COMPANY  INC. 


STEAM  POWEE  PLANT 
AUXILIAKIES  AND  ACCESSOETES 


TERRELL  CROFT,  EDITOR 

CONSULTING    ENGINEER.       DIRECTING    ENGINEER,    TERRELL    CROFT   ENGINEERING    CO. 
MEMBER   OF   THE    AMERICAN   SOCIETY    OF   MECHANICAL   ENGINEERS. 
MEMBER   OF  THE   AMERICAN  INSTITUTE   OF  ELECTRICAL  ENGINEERS. 

MEMBER   OF  THE   ILLUMINATING  ENGINEERING   SOCIETY. 
MEMBER    OF   THE    AMERICAN   SOCIETY   TESTING   MATERIALS. 


FIRST  EDITION 
FOURTH  IMPRESSION 


McGRAW-HILL  BOOK  COMPANY,  INC. 
NEW  YORK:  370  SEVENTH  AVENUE 

LONDON:  6  &  8  BOUVERIE  ST.,  E.  C.  4 


brary 


GIFT  OF 


COPYRIGHT,  1922,  BY  TERRELL  CROFT 


PRINTED   IN   THE   UNITED   STATES   OF   AMERICA 


LIBRARY 


THE  MAPLE  PRESS  -  YORK  PA 


PREFACE 

Most  of  the  preventable  losses  in  the  engine  rooms  of  steam 
power  plants  occur  in  connection  with  the  auxiliary  equipment. 
Generally  speaking,  there  is  not  a  great  deal  that  the  operating 
engineer  can  do  to  increase  the  efficiencies  of  the  prime  movers 
— the  turbines  or  engines.  It  is  also  a  fact  that,  as  a  rule, 
the  prime  movers  in  a  plant  give  relatively  little  trouble  and 
involve  relatively  little  maintenance  expense.  Most  of  the 
trouble  and  maintenance  expense  is  due  to  the  auxiliaries. 
Thus,  it  follows  that,  in  a  sense,  the  auxiliary  equipment  com- 
prises the  most  important  part  of  that  portion  of  the  power- 
plant  equipment  which  transforms  the  heat  in  the  steam  into 
power. 

Hence,  in  this  book,  it  has  been  the  endeavor  to  give  such 
data  as  will  enable  the  operator  to  select,  and  properly  install, 
auxiliary  equipment  which  will  insure  the  generation  of  power 
at  the  least  cost.  Furthermore — and  quite  as  important — it 
has  been  the  aim  to  provide  the  information  whereby  this  auxil- 
iary equipment  can  be  so  operated  and  maintained  that  its 
preventable  losses  will  be  a  minimum  and  that  its  up-keep 
expense  will  be  as  small  as  is  feasible. 

Drawings  for  all  of  the  411  illustrations  were  made  especially 
for  this  work.  It  has  been  the  endeavor  so  to  design  and 
render  these  pictures  that  they  will  convey  the  desired  infor- 
mation with  a  minimum  of  supplementary  discussion. 

Throughout  the  text,  principles  which  are  presented  are 
explained  with  descriptive  expositions  or  with  worked-out 
arithmetical  examples.  At  the  end  of  each  of  the  13  divisions 
there  are  questions  to  be  answered  and,  where  justified, 
problems  to  be  solved  by  the  reader.  These  questions  and 
problems  are  based  on  the  text  matter  in  the  division  just 
preceding.  If  the  reader  can  answer  the  questions  and 
solve  the  problems,  he  then  must  be  conversant  with  the  sub- 
ject matter  of  the  division.  Detail  solutions  to  all  of  the 
problems  are  printed  in  the  appendix  in  the  back  of  the  book. 


viii  PREFACE 

As  to  the  method  of  treatment:  Pumps  are  first  considered 
because  almost  every  power  plant,  regardless  of  size,  requires 
pumps  of  some  sort,  for  its  operation.  Hence,  there  are 
divisions  on  pump  calculations,  direct-acting  steam  pumps, 
crank-action  pumps,  centrifugal  and  rotary  pumps.  Next 
follows  a  discussion  of  boiler-feeding  apparatus  such  as  boiler- 
feed  pumps  and  their  governors,  injectors,  and  gravity  boiler- 
feeding  devices.  The  problems  of  feed-water  heating  are  then 
treated  in  the  divisions  on  feed-water  heaters  and  economizers. 

Following  this  are  divisions  on  condensers  and  methods  of 
recooling  condensing  water  which,  it  is  believed,  are,  both 
economically  and  practically,  very  thoroughly  treated. 
Finally,  the  divisions  on  steam  piping,  live-  and  exhaust- 
steam  separators,  and  steam  traps  explain  how  these  elements 
should  be  selected,  installed,  and  maintained.  They  also 
present  solutions  to  the  problems  of  preventing  losses  from 
and  in  steam  pipes. 

With  this,  as  with  other  books  which  have  been  prepared  by 
the  author,  it  is  the  sincere  desire  to  render  it  of  maximum 
usefulness  to  the  reader.  It  is  the  intention  to  improve  the 
book  each  time  it  is  revised  and  to  enlarge  it  as  conditions  may 
demand.  If  these  things  are  to  be  accomplished  most  effec- 
tively, it  is  essential  that  the  readers  cooperate  with  us.  This 
they  may  do  by  advising  the  author  of  alterations  which  they 
feel  it  would  be  advisable  to  make.  Future  revisions  and 
additions  will,  insofar  as  is  feasible,  be  based  on  such  sugges- 
tions and  criticisms  from  the  readers. 

Although  the  proofs  have  been  read  and  checked  very  care- 
fully by  a  number  of  persons,  it  is  possible  that  some  undis- 
covered errors  may  remain.  Readers  will  confer  a  decided 
favor  in  advising  the  author  of  any  such. 

TERRELL  CROFT. 
UNIVERSITY  CITY, 

ST.  Louis  Mo., 
March,  1922. 


ACKNOWLEDGMENTS 

The  author  desires  to  acknowledge  the  assistance  which 
has  been  rendered  by  a  number  of  concerns  and  individuals 
in  the  preparation  of  this  book. 

Considerable  of  the  text  material  appeared  originally  as 
articles  in  certain  trade  and  technical  periodicals  among 
which  are:  Power,  National  Engineer,  Power  Plant  Engineer- 
ing, and  Southern  Engineer. 

The  author  is  particularly  indebted  to  Mr.  H.  H.  Kelley  and 
to  Mr.  F.  A.  Burg,  manager  of  the  condenser  section  of  the 
Westinghouse  Electric  and  Manufacturing  Company  for  their 
contributions  to  the  condenser  division.  Acknowledgment  is 
also  here  given  to  Mr.  F.  F.  Nickel  for  his  able  assistance  in 
the  matter  on  pumps. 

Among  the  manufacturers  who  cooperated  in  supplying 
text  data  and  illustrations  are:  The  Cooling  Tower  Company; 
Worthington  Pump  and  Machinery  Corporation;  Union  Steam 
Pump  Company;  The  Goulds  Manufacturing  Company;  Schutte 
and  Kcerting  Company;  H.  S.  B.  W.-Cochrane  Corporation; 
Green  Fuel  Economizer  Company;  B.  F.  Sturtevant  Company; 
Westinghouse  Electric  and  Manufacturing  Company;  C.  H. 
Wheeler  Manufacturing  Company;  Wheeler  Condenser  and 
Engineering  Company;  Spray  Engineering  Company;  Crane 
Company. 

Special  acknowledgment  is  hereby  accorded  Edmond  Siroky, 
Head  Mechanical  Engineer  of  The  Terrell  Croft  Engineering 
Company,  who  has  been  responsible  for  the  technical  accuracy 
of  the  book. 

Other  acknowledgments  have  been  made  throughout  the 
book.  If  any  has  been  omitted,  it  has  been  through  oversight 
and,  if  brought  to  the  author's  attention,  it  will  be  incorporated 
in  the  next  edition. 

TERRELL  CROFT. 


IX 


CONTENTS 

PAOE 

FRONTISPIECE.   .    .   .   ...    *  •'.-.; ..    *,.,.<  .,J..* iv 

PREFACE »    . ..  , ,..    .    . ....   .  .....   , ',. vii 

ACKNOWLEDGEMENTS > ix 

LIST  OF  SYMBOLS :......, xii 

DIVISION     1. — PUMP  CALCULATION 1 

DIVISION    2. — DIRECT-ACTING  STEAM  PUMPS 39 

DIVISION    3. — CRANK-ACTION  PUMPS.    .  *   . -u   . 75 

DIVISION    4. — CENTRIFUGAL  AND  ROTARY  PUMPS 101 

DIVISION    5. — INJECTION ,..«  ,.  ^ 155 

DIVISION     6. — BOILER-FEEDING  APPARATUS  (Pump  Governors)      .  171 

DIVISION     7. — FEED- WATER  HEATERS 207 

DIVISION    8. — FUEL  ECONOMIZERS.    .    .    .    ;..*;*;.. 251 

DIVISION    9. — CONDENSERS 277 

DIVISION  10. — METHODS  OF  RECOOLING  CONDENSING  WATER  .   .    .  329 

DIVISION  11. — STEAM  PIPING  OF  POWER  PLANTS 363 

DIVISION  12. — LIVE-STEAM  AND  EXHAUST-STEAM  SEPARATORS      .    .  385 

DIVISION  13. — STEAM  TRAPS .    .- .   v  .  >    .    .    .    '.    .  403 

SOLUTIONS  TO  PROBLEMS 415 

INDEX  .                                                                                                       .  425 


xi 


STEAM    POWER    PLANT    AUXILIARIES    AND    ACCESSORIES 
LIST  OF  SYMBOLS 

The  following  list  comprises  practically  all  of  the  symbols  which  are  used 
in  formulas  in  this  book.  Symbols  which  are  not  given  in  this  list  are 
denned  in  the  text  where  they  are  first  used.  When  a  symbol  is  used  with 
a  meaning  different  from  that  below,  the  correct  meaning  is  stated  in  the 
text  where  the  symbol  occurs. 

SECTION 
SYMBOL  MEANING  FIRST  USED 

A  Piston  area,  in  square  inches 21 

Abh  Area  of  boiler  heating  surface,  in  square  feet 192 

Ai  Internal  area  of  pipe,  in  square  inches 440 

Af  Area,  in  square  feet 19 

Cg  Specific  heat  of  combustion-gases 303 

Cw  Specific  heat  of  water 303 

d  Diameter  of  impeller,  in  inches 121 

di  Internal  diameter  of  pipe,  in  inches 19 

dim  Inside  diameter  of  main  pipe,  in  inches 444 

d0  External  pipe-diameter,  in  inches 448 

dp  Piston-diameter,  in  inches 26 

ds  Steam-piston-diameter,  in  inches 28 

D  Density  of  steam,  in  pounds  per  cubic  foot 440 

Di  Density,  in  pounds  per  cubic  inch 21 

Dc  Duty,  in  foot  pounds  per  100  pounds  of  coal 47 

Dh  Duty,  in  foot  pounds  per  1,000,000  B.t.u 49 

D«  Duty,  in  foot  pounds  per  1,000  pounds  of  steam 48 

et  Coefficient  of  linear  expansion 447 

E  Efficiency  in  per  cent 392 

Eh  Hydraulic  efficiency,  in  per  cent 37 

Ei  Indicated  efficiency,  in  per  cent 35 

Em  Mechanical  efficiency,  in  per  cent 41 

Ero  Efficiency  of  motor,  in  per  cent 138 

Ep  Efficiency  of  pump,  in  per  cent 138 

Et  Total  efficiency,  in  per  cent 42 

Et  Thermal  efficiency,  expressed  decimally 321 

Ev  Volumetric  efficiency,  in  per  cent 25 

Evd  Volumetric  efficiency,  expressed  decimally 26 

g  Acceleration  due  to  gravity  in  feet  per  second,  per  second  = 

32.2 7 

H  Heat,  in  B.t.u 49 

H  Total  heat  of  steam,  in  British  thermal  units  per  pound 244 

Hf  Per  cent,  saving  in  heat-content  of  fuel 244 

xii 


LIST  OF  SYMBOLS  xiii 

SECTION 

SYMBOL                                                      MEANING                                                 FIRST  USED 

Heat,  in  B.t.u.  given  up  by  the  steam  per  hour 348 

Latent  heat  of  vaporization  of  steam 189 

Current,  in  amperes 138 

Condensation,  pounds  per  hour  per  square  foot  of  pipe  surface  .  499 

A  constant 107 

A  constant 406 

Linear  expansion  of  pipe,  in  inches 447 

Length  of  stroke,  in  inches 21 

Minimum  pipe  length  required  for  bend 448 

Pipe-length,  in  inches,  having  resistance  equivalent  to  one 

90-deg.  elbow 446 

Length,  in  feet 39 

Height,  infect 268 

Static  head,  in  feet 5 

Friction  head,  in  feet,  due  to  pump  passages  and  valves 9 

Friction  head,  in  feet,  due  to  pipe  bends .  9 

Fricton  head,  in  feet,  due  to  inlet  flow 9 

Friction  head,  in  feet,  due  to  straight  pipe .  9 

Total  friction  head,  in  feet 9 

Friction  head,  in  feet,  due  to  valves  in  piping 9 

Measured  head,  in  feet,  due  to  delivery  lift, 11 

Measured  head,  in  feet,  due  to  suction  lift 11 

Head,  in  feet,  due  to  back  pressure  on  delivery-pipe  outlet. . .  11 

Total  measured  head,  in  feet 11 

Total  head,  in  feet 12 

Useful  head,  in  feet 34 

Velocity  head,  in  feet 7 

Length  of  pipe-line,  in  feet 448 

Piston-travel,  in  feet  per  minute 26 

Pipe-length,  in  inches  having  resistance  equivalent  to  one 

globe  valve 445 

Width  of  belt,  in  inches 145 

Relative  humidity  of  the  air  expressed  decimally 398 

Revolutions  per  minute 118 

Number  of  strokes  per  minute 21 

Pressure,  in  pounds  per  square  inch 5 

Absolute  pressure,  in  pounds  per  square  inch 189 

Driving  horse  power 41 

Boiler  horse  power 229 

Discharge  pressure,  in  pounds  per  square  inch 49 

Hydrostatic  pressure  head,  in  pounds  per  square  inch 49 

3.1416 19 

Vacuum,  in  inches  of  mercury 322 

Barometer  reading,  in  inches  of  mercury 327 

Intake  pressure,  in  pounds  per  square  inch 49 


xiv  LIST  OF  SYMBOLS 

SECTION 

SYMBOL  MEANING                                                 FIRST  USED 

Pm      Mean  effective  pressure,  in  pounds  per  square  inch 322 

Ps       Steam  pressure,  in  pounds  per  square  inch 28 

Puhp    Useful  hydraulic  horse  power 34 

Pv       Vapor  pressure,  in  inches  of  mercury 398 

Pw       Total  head-pressure,  in  pounds  per  square  inch 28 

Pwhp    Actual  hydraulic  horse  power 35 

T        Absolute  temperature,  on  Fahrenheit  scale 321 

Tf       Temperature,  in  degrees  Fahrenheit 244 

Tf       Temperature  change,  in  degrees  Fahrenheit 447 

Tfa  Average  temperature,  in  degrees  Fahrenheit,  of  water  leaving 

cooling-tower 419 

T/a     Temperature  of  air,  in  degrees  Fahrenheit 452 

T fe      Temperature  of  condensate,  in  degrees  Fahrenheit 344 

Tfd  Final  temperature  of  condensed  steam,  in  degrees  Fahrenheit  189 

Tfd  Dry-bulb-thermometer  temperature,  in  degrees  Fahrenheit  .  .   406 

Tfa     Loss  of  gas  temperature,  in  degrees  Fahrenheit 303 

Tfi  Temperature  of  intake  water  to  injector,  in  degrees  Fahrenheit  189 

Tf8      Temperature  of  steam  used  for  heating  feed-water 266 

Tf,      Temperature  of  steam,  in  degrees  Fahrenheit 189 

Tfw  Wet-bulb-thermometer  temperature,  in  degrees  Fahrenheit  .  .    392 

T/w     Temperature  gain,  water,  in  degrees  Fahrenheit 303 

TfV     Temperature,  of  feed-water  in  degrees  Fahrenheit 309 

T'fu  Temperature  of  feed  water,  in  degrees  Fahrenheit,  at  exit  of 

economizer 309 

U  Coefficient  of  heat  transfer  in  B.t.u.  per  hour,  per  degree 

Fahrenheit  temperature  difference 277 

v          Velocity,  in  feet  per  second 7 

V        Volts 138 

V        Volume  of  condenser,  in  cubic  feet 342 

Va       Volume,  in  cubic  feet  per  minute 19 

VCf     Displacement,  in  cubic  feet  per  minute 21 

V0m     Quantity  of  water,  in  gallons  per  minute 19 

vm       Velocity,  in  feet  per  minute 19 

W       Weight  of  liquid,  in  pounds 31 

W       Weight,  in  pounds  per  minute 21 

Wc      Weight  of  condensation,  in  pounds  per  hour 452 

Wc      Weight  of  coal,  in  pounds 47 

W/  Weight  of  feed-water  entering  heater,  in  pounds  per  hour  .  .  .   262 

WF      Weight  of  feed  water  leaving  heater  in  pounds  per  hour 266 

Wff      Weight  of  gas,  in  pounds,  per  pound  of  coal  burned 303 

Wi      Weight  of  moisture  in  steam,  in  pounds 476 

W,      Steam  rate,  in  pounds  per  hour 262 

W«      Weight  of  steam,  in  pounds 48 

W*w    Pounds  of  water  pumped  per  pound  of  steam 189 

Wu     Useful  work,  in  foot  pounds ,  .     31 


LIST  OF  SYMBOLS  xv 

SECTION 
SYMBOL  MEANING  FIRST  USED 

Ww  Weight  of  water,  in  pounds 189 

Ww  Weight  of  water  evaporated,  per  pound  of  coal  burned,  in 

pounds 303 

Ww  Water  rate,  in  pounds  per  hour 344 

Ww  Weight  of  water  evaporated,  in  pounds  per  square  foot  per  hour  398 

Wwh  Weight  of  water  per  boiler  horse  power  per  hour 229 

X  Slip,  in  per  cent 24 

X  Saving,  in  per  cent 309 

X  Quality  of  steam 189 

X  Ratio..  ....    .   303 


STEAM  POWER  PLANT 
AUXILIARIES  AND  ACCESSORIES 


•D- 
-Discharge 

Valves  '       //^^7= 


•Discharge 
Pipe 


DIVISION  1 
PUMP  CALCULATIONS 

1.  The  Height  To  Which  Water  May  Be  Drawn  By  Pump- 
Suction  depends  principally:  (1)  Upon  the  condition  of  the 
pump  as  regards  the 
tightness  of  its  valves, 
piston-  or  plunger- 
packing  and  piston- 
rod  packing  .  (2) 
Upon  the  water-fric- 
tion in  the  suction 
pipe  (Sec.  8)  and  fit- 
tings. (3)  Upon  the 
temperature  of  the 
water.  (4)  Upon  the 
altitude  above  sea-level. 


\     Travel 


-Pump  Barrel 

Piston 
.--Packing 


P-Moving 

•Element  or 

Piston 

.-Suction 
•     Pipe 


Sue  ft 
Valves 


NOTE.  —  The  practical 
maximum  suction-lift  is 
about  22  feet. 

EXPLANATI  ON.  -  A  t  - 

mospheric  pressure  at 
sea-level  is  about  14.7 
Ib.  per  sq.  in.,  absolute. 
A  2.31-ft.  height  of  water- 
column  is  the  equivalent 
of  1  Ib.  per  sq.  in.  pres- 
sure. On  this  basis  the 
theoretical  suction-lift  at 
sea-level  is  14.7  X  2.31 
=  34  ft.,  nearly.  But 


m*&:-j-~K<< 

si 

^;%-ii 

^^^w^ 

'.".  •  "Piston  Roof  Connecting  ^ 
'.  »•*.  Pump  Piston  With  vj 
'.     ,'  Steam  or  Driving  '•//, 
•    .'•-..  Piston   .  ,--'<v 
•  •  •   '                 .  .  '  -  •  \v 

:•;••/••"• 

y^'v6 

•>:o'.'»  • 

^ 

:V;?V;|| 

;^x 

•  •  .  \y 

•••  :  ;^,.' 

§si 

;^' 

*'<3  •'."•"• 

;•'   •  .v  '  -.  :  .-  ^< 

^syv^s^- 


=---Foof-  Valve-— 


FIG.  1  .  —  How  A  Double-Acting  Suction  Pump  Operates. 


..fi 
in    actual   practice  a  lift 

of   22   ft.   under  sea-level  atmospheric  pressure  is,  due  to  unavoidable 
leakage,  friction  and  vaporization  (Sees.  5,  8,  and  10),  seldom  exceeded. 

1 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


Therefore  (Table  2),  a  pump  lifting  22  ft.  at  sea-level  would,  at  1  mile 
above  sea-level,  where  the  atmospheric  pressure  is  about  12.02  Ib.  per 
sq.  in.,  give  a  lift  of  12.02  X  22  -=-  14.7  =  17.9  ft. 

NOTE. — The  net  suction-lift  of  a  reciprocating  pump  is  the  vertical 
distance,  Lhs  (Fig.  1),  from  the  level  of  the  water  in  the  well,  or  other 
source  of  suction-supply,  to  the  level  of  the  discharge-valve  seats.  The 
total  suction-lift  comprises  the  net  lift  and  the  friction  head  (Sec.  6)  due 
to  water-friction. 

2.  Table  Showing  Practical  Pump  Suction  Lifts  At  Various 
Altitudes. — Ordinary  atmospheric  temperature  is  assumed. 
(Goulds  Catalogue.) 


Altitude  above  sea-level 

Barometric  pressure 

Practical 

suction 

Miles 

Feet 

Pounds 
per  sq.  in. 

Head  in  ft. 
of  water 

lift  of 
pumps, 
feet 

Sea-level 

Sea-level 

14.70 

33.95 

22 

H 

1320 

14.02 

32.38 

21 

y* 

2640 

13.33 

30.79 

20 

H 

3960 

12.66 

29.24 

18 

i 

5280 

12.02 

27.76 

17 

iK 

6600 

11.42 

26.38 

16 

iK 

7920 

10.88 

25.13 

15 

2 

10560 

9.88 

22.82 

14 

3.  In  The  Pumping  Of  Hot  Water  the  tendency  of  the  water 

to  vaporize  under  different  de- 
grees of  absolute  pressure  must 
be  considered.  As  the  suction- 
lift  of  a  pump  increases  (Fig. 
2),  the  maximum  temperature 
of  the  water  that  can  be  pumped 
decreases.  Generally,  it  will  be 
found  practically  impossible  to 
lift  water  at  a  temperature 
above  150  deg.  fahr.  Hence, 
where  a  boiler  feed-pump  (Fig. 
3),  receives  its  suction  supply 

from  an  open  feed-water  heater,  the  water  must  flow  to  the 

pump  under  (Sec.  4)  a  static  head. 


Intake  Heaoi 


inl  Suction 


12   1*    20  24  28  32 
Lift  in  Feet 


FIG.  2. — Diagram  Showing  Practi- 
cal Intake  Pressures  At  Different 
Temperatures;  Also  The  Theoretical 
Water  Lift  Of  A  Pump.  Calculations 
Are  For  Sea  Level. 


SEC.  4] 


PUMP  CALCULATIONS 


3 


4.  The  Static  Head  Of  A  Fluid  Column,  as  a  column  of  water 
(Lh2,  Fig.  4)  in  a  standpipe,  is  the  vertical  distance  between 
the  base  and  the  top  surface  of  the  column.  It  is  understood 
to  mean  the  pressure  which  the  column  imposes  on  the  plane 
which  is  taken  as  a  base.  Thus,  a  30-ft.  static  head  means  the 


CoM  Water  Inlet--- 


Pump  Exhaust-Pipe 
Connection- 


FIG.  3. — Boiler  Feed-Pump  Taking  Water-Supply  From  Open  Feed-Water  Heater. 

pressure,  per  unit  of  base  area,  which  is  due  to  the  weight  of  a 
fluid  column  30  ft.  high. 

EXPLANATION.— LAI  (Fig.  4)  is  the  static  head  of  the  column  of  water 
above  the  plane  AB,  while  Lhz  is  the  static  head  above  the  plane  XY. 
A  column  of  water  1  in.  square  and  1  ft.  high  weighs,  approximately, 
0.433  Ib.  Hence,  the  static  heads,  Lh\  and  LM,  may  be  readily  trans- 
lated into  terms  of  pressure.  Thus,  if  L/u  =  40  ft.,  then  the  pressure  on 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


Stcrncf-Pipe---) 


AB  =  0.433  X  40  =  17.32    Ib.    per  sq.   in.     If  LM  =  50  ft.,    then   the 
pressure  on  XY  will  be:  0.433  X  50  =  21.65  Ib.  per  sq.  in. 

NOTE. — THE  INLET  STATIC  HEADS  (!NLET  PRESSURES)  FOR  BOILER 
FEED-PUMPS  (Lh,  Fig.  3)  drawing  water  from  open  feed-water  heaters 
should,  in  order  to  secure  satisfactory  service,  be  from  about  1.5  ft.  for 
water  at  165  deg.  fahr.  to  about  11.5  ft.  for 
water  (Fig.  2)  at  210  deg.  fahr.  The  pressures 
due  to  these  heads  are  necessary  to  counteract 
the  tendency  of  pumps  to  become  steam-bound, 
or  filled  with  vapor  from  heated  water.  When 
a  reciprocating  pump  is  in  this  condition,  the 
piston  traverses  the  cylinder  without  producing 
a  discharge.  The  vapor  in  each  end  of  the  cyl- 
inder is  compressed  during  one  stroke  and  re- 
expands  during  the  opposite  stroke,  while  the 
boiler-pressure  above  the  delivery  valves  holds 
them  seated. 


HD  ILL  M;:'-:!II 

i=l^ 

L|"    Lhl 

A 
X 

l^llill 

\Y 

FIG.  4. — Illustrating  Static 
Head  Of  A  Liquid. 


5.  The  Head  Or  Pressure  Due  To  A 
Column  Of  Water  May  Be  Converted 
Into  Equivalent  Terms  Of  Unit  Pressure  by  the  following 
formula : 

(1)         P  =  0.433Lft  =  2~o7  (pounds  per  square  inch) 

Wherein  P  =  pressure,   in  pounds  per  square  inch.     Lh  = 
static  head,  in  feet. 

EXAMPLE. — A  direct-acting  steam  pump  (Fig.  1)  is  discharging  into 
an  open  tank.  What  is  the  pressure,  due  to  the  discharge  head,  on  the 
discharge-end  of  the  pump-plunger  if  the  vertical  distance  from  the 
horizontal  axis  at  the  pump-cylinder  to  the  level  of  the  water  in  the  tank 
is  25  feet? 

SOLUTION.— By  For.  (1)  P  =  LA/2.31  =  25  4-  2.31  =  10.8  Ib.  per 
sq.  in. 

6.  A  Pump  Must  Overcome  Certain  Resistances  And 
Pressures  in  delivering  water  or  other  liquids.  The  following 
must  be  considered  in  calculations: 

(1)  Velocity  head  or  velocity  pressure,  which  is  the  head  or 
pressure  required  to  set  the  liquid  in  motion  and  give  it  the 
velocity  which  it  will  have  at  the  final  stage  of  its  movement. 

(2)  Friction  head  or  friction  pressure,  which  is  the  resistance 
head  or  pressure  required  to  overcome  the  resistance  due  to 


SEC.  7]  PUMP  CALCULATIONS  5 

the  friction  between  the  liquid  and  the  surfaces  of  the  pipes, 
fittings,  valves  and  pump-passages  through  which  it  flows. 

(3)  Measured  head  or  measured  pressure,  which  is  the  vertical 
height,  or  the  equivalent  pressure  due  to  this  height,  from  a 
lower  to  a  higher  plane  in  the  pumping  system.  The  lower 
plane  may  be  the  surface  of  a  cooling  pond.  The  higher  plane 
may  be  the  center  of  the  mouth  of  the  discharge  pipe  which 
conveys  the  water  into  a  tank. 

NOTE. — THE  DYNAMIC  HEAD  OR  PRESSURE  is  the  sum  of  the  velocity- 
head  and  friction-head. 

7.  The  Velocity  Of  A  Liquid  In  A  Pipe  Must  Be  Produced 
By  Pressure.  The  pressure  may  be  thought  of  as  the  pressure 
which  is  produced  (Sec.  5)  by  a  vertical  column  of  the  liquid. 
If  friction  and  all  other  resistances  are  neglected,  the  velocity 
produced  by  a  certain  head  will  be  equivalent  to  the  velocity 
attained  by  a  falling  body  which  descends  a  distance  equal 
to  the  head.  See  also  Div.  4.  It  can  be  shown  that: 

(2)  v  =  -\/2g  Lhv  (feet  per  second) 
Wherein  v  —  velocity,  in  feet  per  second,     g  =  acceleration 
due  to  gravity,  in  feet  per  second  per  second  =  32.2  approxi- 
mately.    Lhv  =  head  necessary  to  produce  the  velocity,  in 
feet. 

If  the  velocity  is  known,  the  head  to  which  it  is  due  may 
be  found  by  the  above  formula  rearranged : 

(3)  Lhv  =  |^  (feet) 

NOTE. — As  the  velocity  is  often  small,  the  hydraulic  head  necessary 
to  produce  it  will  be  small.  It  is,  therefore,  often  neglected.  See 
following  sections  and  examples. 

EXAMPLE. — What  velocity  will  result  from  a  head  of  50  ft.  of  water 
when  all  the  head  is  available  for  imparting  velocity  to  the  water? 


SOLUTION.— By     For.     (2)      v  =    V2gLhv  =  \/2  X  32.2  X  50  =  56.7 
ft.  per  sec. 

EXAMPLE. — What  velocity  head  must  a  pump  produce  if  it  is  to  dis- 
charge a  liquid  at  a  velocity  of  10  ft.  per  sec.  ?  SOLUTION. — By  For.  (3) : 
Lhv  =  v*/2g  =  (10)2  -  (2  X  32.2)  =  1.58 /*. 

8.  The  Friction-Head  On  A  Pump  may  be  necessary  for 
overcoming  the  following  resistances:  (1)  The  friction  (Tables 
14  and  15)  due  to  the  flow  of  a  liquid  through  straight  piper. 


6 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


(2)  The  friction  of  the  liquid  entering  (Figs.  5,  6,  and  7)  the 
suction  or  inlet  pipe.  (3)  The  friction  due  to  the  flow  through 
the  pump-valves  and  passages  within  the  pump.  (4)  The  fric- 


Suction 
Pipe 


Square  Orif/'ce -' 

FIG.  5. — Pump  Suction- 
Pipe  With  Square  En- 
trance Orifice. 


Coupling 


Fio.    6. — Pump    Suction- 
Pipe  With  Strainer. 


FIG.  7. — Funneled  End  Of 
Pump  Suction-Pipe. 


tion  due  to  the  flow  (Figs.  8,  9,  and  10)  through  pipe-fittings; 
this  resistance  is  caused  by  the  change  of  direction  of  the 
flow,  and  by  the  roughness  of  the  fittings.  (5)  The  friction  due 
to  flow  through  valves  in  the  piping;  with  gate  valves  this  resist- 
ance is  negligible. 


Direct/on  of  Flow-. 


Direction  of  Fhw-.. 


'•Couplings 


FIG.  8. — Turn  In  Pump 
Piping  Made  With  Long- 
Radius  Bend. 


FIG.  9. — Turn  In  Pump- 
Piping  Made  With  Elbow 
Having  A  Radius  Equal 
To  Pipe-Diameter. 


FIG.  10. — Sharp  Turn 
In  Pump  Piping  Made 
With  Plugged  Tee. 


NOTE. — The  head  due  to  friction  of  the  water  entering  the  suction 
pipe  is  called  the  entrance-head. 

9.  The  Total  Friction-Head  On  A  Pump  is  the  sum  of  the 

resistances  enumerated  in  Sec.  8.     It  may  be  expressed  by  the 
following  formula: 


SEC.  10] 


PUMP  CALCULATIONS 


(4)  Lh/T  =  Lhfp  +  Lhfi  +  Lhff  +  LhfV  +  Lhfc          (feet) 

Wherein  LA/T  =  the  total  friction  head,  in  feet.  LhfP  = 
the  friction-head,  in  feet,  due  to  flow  through  straight  runs 
of  suction  and  discharge  piping.  Lhfi  =  the  friction-head, 
in  feet,  due  to  the  inlet  flow.  Lhff  =  the  friction-head,  in 
feet,  due  to  pipe-fittings  which  change  the  direction  of  the 
flow.  Lhfv  =  the  friction-head,  in  feet,  due  to  valves  in  the 
piping.  Lh/c  =  friction-head,  in  feet,  due  to  flow  through 
passages  and  valves  in  pumps. 

10.  The    Measured   Heads   In  Pump   Operation,   LHh   in 
Fig.    11    (see  Sec.  6  for  definition  of  Measured  Head)  com- 


Dlschctr0e  Level 


.-•5 


K'7,  ^  Sucthn  Lev-e/^. ..'.  •  ; .      .• ;   . 


FIG.   11. — Illustrating  Useful  Pump- Work. 


prise  the  following:  (1)  The  suction-lift  of  the  water.  The 
height  of  this  lift  is  (Lhms,  Fig.  11)  from  the  level  of  the  dis- 
charge valve  seat  to  the  surface  of  the  suction-water.  (2) 
The  delivery-lift  of  the  water.  This  (Lhmdj  Fig.  11)  is  mea- 
sured from  the  level  of  the  seat  of  the  discharge  valve  to  the 
center  of  the  outlet  orifice  of  the  delivery  pipe,  where  the  water 
issues  horizontally  from  the  pipe  and  falls  by  gravity.  Or 


8 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


(Fig.  11)  it  is  measured  from  the  level  of  the  discharge  valve 
seat  to  the  level  of  the  water  above  the  discharge  orifice, 
where  the  outlet  end  of  the  pipe  is  submerged.  (3)  The 
head  due  to  pressure,  above  atmospheric  pressure,  on  the  liquid 
in  the  vessel  into  which  the  delivery-pipe  discharges.  If  water 
is  being  delivered  to  a  boiler,  this  head  is  equivalent  to  the 
steam-gage  pressure  in  the  boiler. 

NOTE. — THE   SUCTION  LIFT  OF  A  CENTRIFUGAL  PUMP  is  measured 
from  the  level  of  the  water  in  the  well  to  the  center  of  the  impeller. 


Air  Pockef'Due  To  Hummock  ]n  Pipe  Line--, 


vr'v •*---•• 

'-iS'^Live  Of  Proper  DeclMly  Xr^ .  =\  'J^r,'^ 
'. •' .-  •_•.-'  ,  Toward  ffes«m>r---T-v £  ;  ... .'  ^J.':-'  -  l^ 

'"    -~-      ' 


FIG.   12. — An  Imperfectly  Laid  Suction-Line. 

NOTE. — When  suction  pipes  are  laid  underground,  in  trenches,  care 
should  be  exercised  to  run  them  in  a  slightly  declining  straight  line 
toward  the  source  of  supply.  High  places  or  hummocks  (Fig.  12)  in 


-Gate  VdtviJ \- .; ^  TL: N  = .  =  ', "-,- , M'-^ ^  "- '•'' ~~. '",.' : 


^^T^v^^s^jssw.srt'g^gSjgCT! 

FIG.   13. — Suction  Well  Supplied  Through  Intake  Pipe. 


the  suction  line  afford  pockets  for  the  accumulation  of  air.  Such  air- 
pockets  reduce  the  effective  area  of  the  pipe  and  cut  down  the  water 
supply  to  the  pump. 


SEC.  11]  PUMP  CALCULATIONS  9 

NOTE. — When  the  distance  from  a  pump  to  a  natural  source  of  suction- 
supply,  as  a  pond,  lake,  or  stream,  exceeds  about  100  ft.,  it  is  advisable 
(Fig.  13)  to  sink  a  suction-well  close  to  the  pump.  The  intake-pipe 
should  then  incline  toward  the  well. 


11.  The  Total  Measured  Head  On  A  Pump  is  the  sum  of 

the  measured  heads  enumerated  in  Sec.  10.  It  may  be  ex- 
pressed by  the  following  formula: 

(5)  LhmT    =   Lhms    -f-  Lhmd   +  Lhmp  (feet) 

Wherein  LhmT  =  the  total  measured  head,  in  feet.  Lhms  = 
the  measured  head,  in  feet,  due  to  suction-lift.  Lhmd  =  the 
measured  head,  in  feet,  due  to  delivery-lift.  LhmP  =  the 
measured  head,  in  feet,  due  to  steam,  compressed  air,  or  other 
fluid  pressure  in  the  vessel  into  which  the  delivery-pipe  dis- 
charges. If  the  delivery  pipe  discharges  freely  into  the 
atmosphere,  then  Lhmp  is  zero. 

12.  The  Total  Head  On  A  Pump  is  the  sum  of:  the  velocity- 
head   (Sec.   7),   the  total  friction-head   (Sec.   9)    and   the  total 
measured-head  (Sec.  11).     It  may  be  expressed  by  the  following 
formula: 

(6)  LhT  =  Lhv  +  LhfT  +  LhmT  (feet) 

Wherein  LhT  =  the  total  head,  in  feet.  Lhv  =  the  velocity- 
head,  in  feet.  LhfT  =  the  total  friction-head,  in  feet.  LhmT  = 
the  total  measured-head,  in  feet. 

13.  The  Friction  Of  Water  In  Straight  Pipes  Is  Difficult 
To  Determine  Definitely  In  All  Cases. — The  smoothness  of 
the  pipe-surface,  the  length  of  time  the  piping  has  been  in 
service,  the  size  of  the  pipe,  and  the  nature  of  the  substances 
with  which  it  may  be  scaled  or  coated  internally,  are  the 
principal    determining    factors.     These    factors    may    vary 
widely,  in  individual  cases,  from  working  standards  which  are 
based  upon  experimental  data. 

NOTE. — The  data  given  herein  (Tables  14  and  15)  are  for  new  pipe. 
When  the  pipe  is  very  rough,  or  is  old  and  rough,  the  actual  values  may 
be  greater  than  those  shown.  In  such  cases,  the  resistance  due  to 
friction  can  be  determined  only  by  tests. 


10 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


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SEC.  14] 


PUMP  CALCULATIONS 


11 


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12 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


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SEC.  15] 


PUMP  CALCULATIONS 


13 


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14 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


EXAMPLE. — A  pump  (Fig.  14)  draws  1050  gal.  of  water  per  min.  from 
a  pond  and  delivers  it,  straightaway,  through  1,020  ft.  of  new  6-in.  steel 
pipe.  What  pressure  is  required  to  force  the  water  against  the  frictional 
resistance  of  the  pipe?  What  additional  pressure  is  required  to  impart 
the  necessary  velocity  of  flow  in  the  delivery  pipe? 

SOLUTION. — By  Table  1*4  the  friction-head  per  100  ft.  of  6-in.  pipe  = 
9.5  ft.  Hence,  the  friction-head  for  the  given  length  of  6-in.  pipe  = 


Centrifugal  Pump 
.-•Priming  Ejector 

h 


6~- In.  Delivery  P/pe--, 
-1020  Feet- 
3 


FIG.   14.— Illustrating  Delivery  Through  Straight  Run  Of  Pipe. 


1,020  ^  100  X  9.5  =  96.9  ft.     By  For.  (1),  P  =  0.433  Lh  =  0.433  X 
96.9  =  42  Z6.  per  sq.  in. 

By  Table  14,  the  velocity  =  11.9/4.  per  sec.  Hence,  by  For.  (3)  Lhv  = 
v*/2g  =  11.92  -M2  X  32.2)  =  2.2ft.  By  For.  (1)  P  =  0.433  Lh  = 
0.433  X  2.2  =  0.954  Ib.  per  sq.  in.  This  velocity-head  is  so  small  that  it 
could,  in  practice,  be  neglected  without  appreciable  error. 

16.  The   Heads  Necessary  To  Overcome  The  Frictional 
Resistance   To   Water-Flow   Through   Fittings   And   Valves 
depend  principally  upon  the  ages  and  the  sizes  of  the  fittings 
and  valves,  and  upon  the  relative  smoothness  of  their  sur- 
faces.    Approximate  values  are  given  in  Table  18. 

17.  The  Frictional  Resistance  Offered  By  The  Internal 
Passages  And  Valves  Of  A  Pump  is  very  small.     Often  it  is 
equivalent  to  a  head  loss  of  only  1  ft.     The  maximum  seldom 
exceeds  3  ft. 


SEC.  18] 


PUMP  CALCULATIONS 


15 


18.  Table  Showing  Approximate  Length,  In  Feet,  Of 
Straight,  Clean  Wrought  Iron  Or  Steel  Pipe  In  Which  The 
Frictional-Resistance  Is  Equivalent  To  That  In  The  Fittings 
Listed. 


Size  of  pipe  and  fittings,  in 
inches  

M 

H 

1 

IK 

1H 

2 

2K 

3 

4 

5 

6 

Elbows,  90  deg.  (Fig.  9)  

5 

6 

6 

8 

8 

8 

11 

15 

16 

18 

18 

2 

3 

3 

4 

4 

4 

6 

g 

g 

g 

10 

Long  radius  bends  (Fig.  8)  ... 

2 

2 

3 

3 

3 

4 

6 

8 

9 

9 

10 

Sharp  bends  (Fig.  10)  

10 

12 

12 

16 

16 

16 

22 

30 

32 

36 

36 

Return  bends  

10 

12 

12 

16 

16 

16 

22 

30 

32 

36 

36 

Globe  valves  

5 

6 

6 

8 

8 

8 

11 

15 

16 

18 

18 

Strainer      or     footvalve     at 
entrance     to     suction     pipe 
(Fig.  6)  

10 

12 

12 

16 

16 

16 

22 

30 

32 

36 

36 

Square  kept  entrance  to  suc- 
tion pipe  (Fig.  5)  

5 

6 

6 

8 

8 

8 

11 

15 

16 

18 

18 

Funnel  end  entrance  to  suc- 
tion pipe  (Fig.  7)  

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

0 

EXAMPLE. — A  boiler-feed  pump  delivers  45  gal.  of  water  per  min. 
It  lifts  the  water,  by  suction,  through  a  height  (Lh8,  Fig.  15)  of  6  ft. 
The  suction  piping  is  of  2-in.  size.  It  extends  5  ft.  below  the  surface  of 
the  water  in  the  suction-well.  It  runs  horizontally  for  a  distance,  L, 
of  60  ft.  It  makes  two  right-angled  turns,  TI  and  T2,  by  means  of  plugged 
tees,  and  one  right-angled  turn,  E\,  by  means  of  a  90-deg.  elbow.  The 
water  enters  the  suction-pipe  through  an  orifice,  O,  which  is  formed  by 
cutting  the  pipe  squarely  across.  The  delivery  piping  is  of  1.5-in.  size. 
It  contains  140  ft.  of  straight  pipe,  three  90-deg.  elbows,  Et,  E3  and  E*, 
one  globe  check-valve,  Vi,  and  one  globe  stop-valve,  F2.  The  vertical 
height,  Lhdj  of  the  discharge-lift  is  35  ft.  The  boiler  steam-pressure  is 
110  Ib.  per  sq.  in.  What  is  the  total  head  on  the  pump? 

SOLUTION. — By  Table  14  the  velocity  in  the  straight  runs  of  suction- 
piping  =  4.6  ft.  per  sec.  Also,  the  velocity  in  the  straight  runs  of 
delivery-piping  =  7.08  ft.  per  sec.  Hence,  by  For.  (3),  Lhv  =  vz/2g  = 
(7.08)2  -5-  (2  X  32.2)  =  0.778  ft.  =  velocity-head. 

The  total  length  of  straight  suction-piping  =  6  +  5  +  60  =  71  ft. 
By  Table  14,  the  friction-head  due  to  the  straight  suction-piping  =  (71  -£• 
100)  X  5.8  =  4.118/2.  Also,  the  friction-head  due  to  the  straight  delivery- 


16 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


piping  =  (140  +  100)  X  16.6  =  23.24  ft.  Hence,  the  friction-head  due 
to  straight  piping  in  the  complete  system  =  4.118  +  23.24  =  27.358  ft. 

By  Table  18,  the  2-in.  straight-pipe  equivalent  of  a  suction-inlet  orifice 
formed  by  a  square-cut  pipe-end  =  8  ft.  Hence,  the  entrance-head,  or 
friction-head  due  to  the  inlet  orifice,  =  (8  -r-  100)  X  5.8  =  0.464  ft. 

By  Table  18,  the  2-in.  straight-pipe  equivalent  of  a  sharp  bend  =  16  ft. 
Hence,  the  friction-head  due  to  the  plugged  tees,  Tl  and  Tt,  =  (16  -5-  100)  X 
5.8  X  2  =  1.856  ft.  The  1.5-in.  and  2-in.  straight-pipe  equivalents 
of  a  9Q-deg.  elbow  =  8  ft.  Hence,  the  friction-head  due  to  the  elbows,  E\, 
Ez,  E3,  and  E*  =  (8  -^  100)  X  5.8  X  4  =  1.85  J  ft.  The  friction-head 
due  to  the  six  turns  in  the  piping  is,  therefore,  1 .856  +  1.856  =  3.712  ft. 

By  Table  18,  the  1.5-w.  straight-pipe  equivalent  of  a  globe-valve  =  8  ft. 
Hence,  the  friction-head  due  to  the  valves  Vi  and  F2  =  (8  -f-  100)  X  5.8  X 
2  =  0.928  ft. 


.-Suction  Inlet 


.-Pump  \ 


Horizontal  Length  of 
Suction  Piping- — „• 


FIG.   15. — Pump-Piping  With  Large  Resistance-Head. 


By  Sec.  17,  assume  loss  due  to  flow  through  passages  and  valves  in 
chamber  =  2  ft. 

By  For.  (4),  the  total  friction-head  =  LhfT  =  Lhfp  -f  Lhfi  -f  Lhff  + 
Lhfv  +  Lhfc  =  27.358  +  0.464  +  3.712  +  0.928  +  2  =  34.462  ft. 

By  transposition  of  For.  (1)  the  static-head  equivalent  of  the  boiler- 
pressure  =  Lh  =  2.31  P  =  2.31  X  110  =  254.1  ft.  Hence,  by 

For.  (5),  the  total  measured-head  =  LhmT  =  Lhms  -\-Lhmd  +  Lhmp  = 
6+35  +254.1  =  295. 1ft. 

By  For.  (6),  the  total  head  on  the  pump  =  LhT  =  Lhr  +  Lh/T  + 
LhmT  =  0.778  +  34.462  +  295.1  =  330.34  ft. 

EXAMPLE. — A  steam  pump  (Fig.  16)  has  a  suction-lift,  L^,  of  8  ft.,  and 
a  discharge-lift,  Lhd,  of  82  ft.  The  suction-piping  is  of  3-in.  size.  It 
contains  75  ft.  of  straight  pipe,  one  long-radius  bend,  B,  and  a  funneled 


SEC.  18] 


PUMP  CALCULATIONS 


17 


inlet-orifice,  F.  The  delivery-piping  is  of  2.5-in.  size.  It  contains 
517  ft.  of  straight  pipe,  two  90-deg.  elbows,  E^  and  E2,  and  one  globe 
valve,  V.  It  is  assumed  that  the  head  necessary  to  impart  velocity  to 
the  water  is  (Note  subjoined  to  Sec.  7)  practically  negligible.  In 
practice  the  velocity  head  is,  usually,  practically  zero.  It  is  also  assumed 
that  a  resistance  =  to  2  ft.  is  offered  to  the  flow  through  the  valves  and 
passages  of  the  pump  itself.  The  pump  discharges  into  an  open  reservoir. 
It  is  capable  of  operating  against  a  total  head  which  is  equivalent  to  a 
pressure  of  85  Ib.  per  sq.  in.  What  is  the  maximum  average-rate, 
in  gals.,  per  min.,  at  which  the  pump  can  deliver  the  water  through  this 
system. 


Pump-, 


Horizontal  Length  of 
'  Suction  Piping*, 


FIG.    16. — Pump-Piping  With  Small  Resistance-Head. 

SOLUTION. — First,  find  the  equivalent  frictional  resistances  of  the 
fittings  in  both  the  suction  piping  and  the  discharge  piping  and  reduce  all 
to  the  basis  of  3-in.  piping  as  explained  below:  By  Table  18,  the  3-in. 
straight-pipe  equivalent  of  the  long-radius  bend,  B,  =8  ft.  Also,  the 
straight-pipe  equivalent  of  the  funneled  inlet-orifice,  F,  =  0.0  ft.  Hence, 
the  frictional  resistance  in  the  complete  3  in.  suction-piping  is  that  which 
would  occur  in  a  straight  run  of  75  +  8  =  83  ft.  of  3-in.  pipe. 

By  Table  18,  the  2.5-in.  straight-pipe  equivalent  of  the  two  90-deg.  elbows, 
Ei  and  E2,  =  (11  X  2)  =  22  ft.  Also,  the  straight-pipe  equivalent  of  the 
globe-valve,  V,  =  11  ft.  Hence,  the  frictional  resistance  in  the  complete 
2.5-w.  delivery-piping  is  that  which  would  occur  in  a  straight  run  of  (517  + 
22  +  11  +  2)  =  552ft.  of  2.5-in.  pipe. 

Now  by  comparing  the  Friction-Head  values  for  "2%-Inch  Pipe" 
and  for  ''  3-Inch  Pipe"  from  Table  14,  it  will  be  found  that,  on  the  average, 
2  3^ -in.  pipe  offers  2.4  times  as  much  frictional  resistance  for  the  same 
flow,  in  gallons  per  minute,  as  does  3-in.  pipe.  Hence,  the  frictional 
resistance  in  the  entire  piping  system  is  equivalent  to  that  which  would  occur 
in  a  straight  run  of:  83  +  (552  X  2.4)  =  1408  ft.  of  3-in.  pipe. 
2 


18  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

By  For.  (5),  the  total  measured  head,  LhmT  =  Lhm*  +  Lhmd  = 
8  +82  =  90ft. 

By  transposition  of  For.  (1),  the  total  static  head  developed  by  the  pump 
which  is  equivalent  to  a  pressure  of  85  Ib.  per  sq.  in.  =  L&  =  2.31P  = 
2.31  X  85  =  196.35  ft. 

Hence,  the  head  which  remains  or  which  is  available  for  overcoming  the 
fractional  resistance  of  the  entire  pumping  system,  that  is,  the  frictional 
resistance  or  head  of  1408  ft.  of  3-in.  pipe  =  196.35  -  90  =  106.35  ft. 
Stating  this  in  friction  head  per  100  ft.  of  straight  pipe:  (106.35  -3-  1408) 
X  100  =  7.55  ft.  friction  head  per  100-/Z.  length  of  3-in.  pipe. 

By  Table  14,  the  flow  corresponding  to  a  friction-head  of  7.72  ft.  per 
100-/2.  length  of  3-in.  pipe  =  150  gal.  per  min.  Hence,  the  flow  corre- 
sponding to  a  friction  head  of  7.55  ft.  per  100-ft.  length  of  3-in.  pipe  is, 
approximately:  (7.55  X  150)  -j-  7.72  =  146.7  gal.  per  min.  =  the  maxi- 
mum average  rate  of  delivery. 

19.  The  Proper  Sizes  For  The  Suction  Or  Discharge  Pipe 
Of  Any  Pump  may  be  computed  with  the  following  formula, 
the  derivation  of  which  is  given  below: 


(7)  di  =  4.95  (inches) 

\     ^m 

Wherein,  d,  =  actual  internal  diameter  of  suction — or  dis- 
charge— pipe,  in  inches.  V gm  =  amount  of  water  to  be  pumped, 
in  gallons  per  minute.  vm  =  average  velocity  of  flow,  in 
feet  per  minute. 

NOTE. — Transposing  the  above  there  results: 

94  f\V 

(8)  vm  =  ^^  (feet  per  min.) 

(9)  Vgm  =  ^j  =  0.004  di*vm  (gallons  per  min.) 

DERIVATION. — Since,  Fc/,  the  amount  of  water  to  be  pumped  in  cubic 
feet  per  minute  =  (A  /,  the  cross-sectional  area  of  the  suction  or  discharge 
pipe,  in  square  feet)  X  (vm,  the  allowable  velocity  of  flow,  in  feet  per  minute), 
it  follows  that: 

(10)  Vef  =  AfVm  =  l  (^)  2  vm  =  ^-ro  (cu.  ft.  per  min.) 
Also,  since  1  cu.  ft.  =  7.48  gal: 

(11)  Vc/  =  £||  (cu.  ft.  per  min.) 

Now,  equating  (10)  and  (11): 
Vm 


7.48        576 


SEC.  20] 


PUMP  CALCULATIONS 


19 


Then,  solving  for  d»: 


(13) 


i  =  4.95 


(inches) 


Piston-. 


EXAMPLE.  —  A  simple  direct-acting  steam  pump  is  required  to  deliver 
800  gals,  of  water  per  min.  What  should  be  the  diameter  of  the  suction 
pipe  if  the  allowable  flow  velocity  in  it  (See  Sec.  52)  is  200  ft.  per  min.  ? 
What  should  be  the  diameter  of  the  discharge  pipe  if  the  allowable  flow 
velocity  in  i  is  400  ft.  per  min.?  SOLUTION.  —  For  he  suction  pipe 
by  For.  (7),  di  =  4.95VVgm/vm  =  4.95\/800  -=-  200  =  9.9  in.,  or,  practi- 
cally, a  10-in.  internal-diameter  pipe.  For  the  discharge  pipe:  di  = 
4.95\/800/400  =  6.98  in.  or,  practically,  a  7-in.  internal-diameter  pipe. 

20.  The     Displacement    Of    A 
Reciprocating    Pump    is   the    vol- 
ume   of    space    (Fig.    17)    swept 
through  by  the  piston  or  plunger 
in    a    definite    interval    of   time. 
Assuming    the    pump-cylinder    to 
be  full  of  water  at  the  beginning 
of   each  stroke,  the  displacement 
is  equal  to  the  volume  of  water 
which  is  driven  out  of  the  cylinder 
during  the  given  time-interval. 

NOTE.  —  The  displacement  of  a  pump 
may  be  expressed  in  cubic  feet,  pounds  or 
gallons  per  minute. 

21.  The  Displacement  Of  Any 
Piston     Or    Plunger    Pump    Per 


Discharge 
Outlet^ 


Displacement 
.-of  One  Stroke 


Suction 
Inlet-- 


FIG.  17.  —  showing  volume  of 


Minute  may  be  found  by  the  fol-     pace    Swept    Through   m  One 

J  J 

lowing  formulae: 


, 

Stroke  Of  Piston. 


(14) 
(15) 
(16) 


Vcf   = 

W»  = 


1728 


LAN* 

231 


(cubic  feet  per  min.) 
(pounds  per  min) 
(gallons  per  min) 


Wherein  Vcf  =  displacement  in  cubic  feet  per  minute.  Ww 
=  displacement  in  pounds  per  minute.  Vgm  =  displacement 
in  gallons  per  minute.  L  =  length  of  stroke  in  inches.  A  = 
effective  area  of  piston  or  plunger,  in  square  inches.  N8  = 
number  of  strokes  per  minute.  Di  =  density  of  liquid  to  be 
pumped  in  pounds  per  cubic  inch. 


20 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


NOTE. — THE  EFFECTIVE  PLUNGER  OR  PISTON  AREA  of  an  outside- 
end-packed  plunger  (Fig.  18)  in  a  direct-acting  steam-pump  is  the 
cross-sectional  area  of  the  plunger.  Of  a  center-packed  (Fig.  19)  or 
inside-packed  (Fig.  20)  plunger,  or  of  a  piston  (Fig.  17),  it  is  the  cross- 


.-P/'sfon  Plunger-Pool 
Cradle-Pools-. 


Discharge  Outlet 

Plunger  - . 
Crossheact  ''• 


FIG.   18.— Water-End    Of    Direct-Acting    Steam-Pump    With    Outside    End-Packed 

Plungers. 

sectional  area  of  the  plunger  or  piston  minus  one-half  the  cross-sectional 
area  of  the  piston-  or  plunger-rod. 

EXAMPLE. — What  is  the  displacement,  in  cubic  feet  per  minute,  of  an 
outside  center-packed  duplex  pump  (Fig.  19),  if:  the  plunger-diameter 
is  18  in.,  the  plunger-rod  diameter  is  3  in.,  the  length  of  stroke  is  24  in., 


Steam  ftsfon- 

.-Wo/ter  Plunger 

Pool 


Glanfr. 


FIG.  19. — Water-End  Of  Direct-Acting  Steam-         FIG.  20. — Pump-Plunger  Inside- 
Pump  With  Outside  Center-Packed  Plungers.  Packed  With  Fibrous  Rings. 

and  each  of  the  2  plungers  (this  being  a  duplex  pump)  makes  50  strokes 
per  min.?  SOLUTION. — By  preceding  Note,  the  effective  plunger  area  = 
(182  X  0.7854)  -  (3*  X  0.7854  •*•  2)  =  250.8  sq.  in.  Now,  substitute 
in  For.  (14):  Vef  =  LANS/1728  =  24  X  250.8  X  50  X  2  +  1728  = 
348.5  cu.  ft.  per  min. 


SEC.  22] 


PUMP  CALCULATIONS 


21 


Discharge  Outlet- 

Valves  V,anol 
V?  in  Act  of 
Closing..  _ 

Suction 
.--Inlet 


.-Discharge  Voitves 


fi'on 


22.  Pump-Slip   is   the   return   of   water,    or  other  liquid, 
through  the  valves  of  a  pump  while  the  valves  (Fig.  21)  are  in 
the  act    of    closing.     It    may    also    occur    by    leakage    past 
the    piston    or    plunger    from    the    discharge-end    to    the 
suction-end  while  the  pump  is  making  a  stroke.     It  is,  there- 
fore, the  difference  between  the  displacement  or  theoretical 
discharge  of  a  pump  and  the  actual  discharge.     It  is  commonly 
expressed  as  a  percentage  of  the  displacement. 

NOTE. — AVERAGE  VALUES  OF  PUMP- 
SLIP,  for  good  pumps,  range  from  3  to 
5  per  cent.  The  slip  of  a  new  pump 
seldom  exceeds  2  per  cent.  Where 
conditions  are  adverse,  the  slip  may 
be  as  great  as  10  or  15  per  cent.  For 
pumps  which  handle  large  volumes  of 
water,  slips  as  low  as  %  per  cent, 
have  been  recorded. 

NOTE. — PUMP-SLIP  MAY  BE  NEGA- 
TIVE. That  is,  the  actual  discharge 
may  be  greater  than  the  theoretical 
discharge.  This  may  occur  if  the 
suction-lift  is  very  low,  and  the 
suction-  and  discharge-lines  run  hori- 
zontally for  considerable  distances. 

The  momentum  of  the  moving  column  (Sec.  65)  may  then  cause  the 
suction  water  to  surge  into  the  cylinder  with  such  force  as  to  produce  a 
considerable  leakage  through  the  discharge  valves  at  the  suction  end. 

23.  Very  High  Piston-Speed  May  Cause  Excessive  Pump- 
Slip. — When  the  piston  reaches  the  end  of  a  stroke,  a  space  of 
time  must  elapse  while  the  open  valves  (Fig.  21)  are  descending 
and  making  firm  contact  with  their  seats.     But,  during  this 
interval,  the  piston  starts  on  the  opposite  stroke.     Some  of  the 
water  that  was  discharged  during  the  preceding  stroke  then 
flows  in  behind  the  piston  through  the  imperfectly  seated  dis- 
charge valves.     Admission  of  a  full  cylinder  of  water  through 
the  suction  valves  is  thus  prevented.     Coincidentally,  some 
of  the  water  ahead  of  the  piston  slips  by  the  suction  valves  and 
passes  back  into  the  suction  chamber. 

24.  The  Percentage  Of  Pump-Slip  May  be  computed  by  the 
following  formula : 

(17)  X  =  -      '—-. (percent,) 

V  cf 


FIG.  21. — How  Pump-Slip  Occurs. 


22  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

Wherein  X  =  per  cent,  of  slip.  Vc/  =  displacement  in  cubic 
feet  per  minute.  V  a  =  actual  discharge  in  cubic  feet  per  minute. 
EXAMPLE.  —  The  displacement  of  a  pump  is  386.85  cu.  ft.  per  min. 
The  pump  delivers  372  A  cu.  ft.  of  water  per  min.  What  is  the  slip? 
SOLUTION.—  By  For.  (17),  X  =  [100(7,.,  -  V  a}}  -=-  Vcf  =  [100  X 
(386.85  -  372.4)]  -i-  386.85  =  3.74  per  cent. 

25.  The  Volumetric  Efficiency  Of  A  Pump  is  the  ratio  of  the 
volume  of  water  actually  delivered  by  the  pump  to  the  dis- 
placement of  the  pump.  It  may  be  computed  by  the  following 
formula  : 


(18)  E,  =          -a  (per  cent.) 

V  cj 

Wherein  Ev  =  the  volumetric  efficiency,  in  per  cent.  Va  = 
the  actual  discharge,  in  cubic  feet  per  minute.  Vcf  =  the 
theoretical  discharge,  or  the  displacement,  in  cubic  feet  per 
minute. 

NOTE.  —  Volumetric  efficiency  and  pump-slip  are  closely  related. 
Thus,  pump-slip  =  1  -  volumetric  efficiency.  Pump-slip  may  vary  from 
0.5  per  cent,  to  15  per  cent.  Hence,  the  volumetric  efficiency  may 
correspondingly  vary  from  99.5  per  cent,  to  about  85  per  cent.  Pump- 
slip  exceeding  2  per  cent,  would  indicate  either  unfavorable  operating 
conditions,  defective  design,  or  a  worn-out  condition  of  the  pump. 

26.  The  Discharge  Of  A  Piston  Or  Plunger  Pump  may  be 
approximately  computed  by  the  following  formula  : 

0.7854dp2LT  _         dp2LTEtd     , 

(19)  Va  =  --  ^-  -E,d=.   i8335-      (cubic  feet  per  mm.) 

Wherein  Va  =  approximate  discharge  capacity,  in  cubic  feet 
per  minute.  dp  =  diameter  of  piston  or  plunger,  in  inches. 
LT  =  the  effective  piston  or  plunger  travel,  in  feet  per  minute. 
Evd  =  the  volumetric  efficiency,  expressed  decimally. 

EXAMPLE.  —  The  plunger  diameter  in  a  direct  acting  duplex  steam  pump 
is  6  in.  The  stroke  is  24  in.  Each  plunger  makes  35  strokes  per  min. 
What  is  the  discharge  when  the  volumetric  efficiency  is  92  per  cent.? 
SOLUTION.  —  The  total  number  of  strokes  per  minute,  this  being  a  duplex 
pump,  =  2  X  35  =  70.  By  For.  (19),  Va  =  dp*LTEvd/  183.35  =  [62  X 
(70  X  24  -T-  12)  X  0.92]  ^  183.35  =  25.3  cu.  ft.  per  min. 

27.  The  Requisite  Diameter  For  The  Water  -End  Of  A  Pump 
Plunger  Or  Piston,  when  the  rates  of  discharge  and  plunger  or 


SEC.  28]  PUMP  CALCULATION  23 

piston    travel    are    given,   may  be  found  by  the  following 
formula  : 

,  /183.35  Va 

(20)  dp  =  (inches) 


Wherein  dp  =  the  diameter  of  plunger  or  piston,  in  inches. 
Va  =  the  actual  discharge,  in  cubic  feet  per  minute.  LT  = 
the  effective  travel  of  the  plunger  or  piston,  in  feet  per  min- 
ute. Evd  =  the  volumetric  efficiency  of  the  pump,  expressed 
decimally. 

EXAMPLE.  —  A  single  direct  acting  steam  pump  is  required  to  discharge 
141  gal.  of  water  per  min.  while  running  90  ft.  of  plunger  travel  per  min. 
If  the  assumed  volumetric  efficiency  of  the  pump  is  97  per  cent.,  what 
should  be  the  diameter  of  the  plunger?  SOLUTION.  —  A  gallon  contains 
231  cu.  in.  By  For.  (20),  dp  = 


V183.35  X  (141  X  231  4-  1728)  +  (90  X  0.97)  =  6.3  in. 

28.  The  Requisite  Steam-Piston  Diameter  For  A  Direct- 
Acting  Steam  Pump  may  be  found  by  the  following  formula  : 


(21)  ds  =        'w  (inches) 

Wherein  ds  =  diameter  of  steam-piston,  in  inches.  A  = 
area  of  water-piston  or  plunger,  in  square  inches.  Pw  =  total 
head-pressure,  in  pounds  per  square  inch.  Ps  =  steam  pres- 
sure, in  pounds  per  square  inch.  The  mechanical  efficiency 
(Sec.  41)  of  the  pump  is  herein  assumed  as  70  per  cent. 

EXAMPLE. — The  requisite  diameter  of  water-piston  for  a  direct-acting 
steam  pump  is  found  to  be  8  in.  The  total  head-pressure  is  200  Ib.  per 
sq.  in.  The  available  steam-pressure  is  80  Ib.  per  sq.  in.  What  should  be 
the  steam-piston  diameter?  SOLUTION. — By  For.  (21)  d8  =  \f\,%APw/Ps 
=  Vl.8  X  82  X  0.7854  X  200  4-  80  =  15  in. 

NOTE. — THE  PLUNGER-  OR  WATER-PISTON-SIZE  FOR  A  DUPLEX 
PUMP  Is  COMPUTED  ON  THE  BASIS  of  one-half  the  total  quantity  of  water 
to  be  delivered,  and  upon  the  rate  of  travel  of  one  piston. 

NOTE. — THE  PISTON-SPEED  OF  A  DIRECT-ACTING  STEAM-PUMP  should 
be  gaged  according  to  the  size,  of  the  pump.  In  large-  and  medium-sized 
pumps  for  general  service,  it  should  not  exceed  about  100  ft.  per  min. 
In  small  pumps,  with  strokes  of  from  about  3  to  9  in.,  the  piston  travel 
should  range  from  about  40  to  75  ft.  per  min. 


24  STEAM  POWER  PLANT  AUXILIARIES  [Dnr.  1 

29.  To   Compute  The  Average  Velocity  Of  Flow  Through 
The  Discharge  Pipe  Of  Any  Reciprocating  Pump,  the  following 
formula  may  be  used.     Slip  is  disregarded. 

79     T 

(22)  vm  =      *r2T  (feet  per  minute) 

Q>i 

Wherein,  vm  =  the  average  velocity  of  flow  through  the  dis- 
charge pipe,  in  feet  per  minute.  dp  =  diameter  of  water  piston, 
in  inches.  d»  =  actual  internal  diameter  of  discharge-pipe,  in 
inches.  LT  =  effective  piston  travel,  in  feet  per  minute;  for  a 
double-acting  pump,  LT  =  feet  which  the  piston  travels  in  a 
minute;  for  a  single-acting  pump,  LT  =  (feet  which  the  piston 
travels  in  a  minute)  +2. 

EXAMPLE.  —  The  water-piston  diameter  in  a  direct-acting  (double- 
acting)  steam  pump  is  3>£  in.  The  discharge-pipe  internal  diam.  is 
1>£  in.  The  piston  travel  is  100  ft.  per  min.  What  is  the  average 
velocity  of  water  flow  in  the  discharge  pipe?  SOLUTION.  —  By  For.  (22): 
vm  =  d2pLT/di2  =  3.5  X  3.5  X  100  4-  (1.5  X  1.5)  =  544  ft.  per  min. 

30.  The  Net  Work  Of  A  Pump  is  the  quantity  of  work 
which  is  theoretically  necessary  to  elevate  the  water  or  other 
liquid  from  the  suction-level  to  the  discharge-level.     That  is, 
it  is  the  work  performed  in  overcoming  the  total  measured 
head,  LhmT  For.  (5). 

31.  The  Net  Work  Performed  By  A  Pump  May  Be  Com- 
puted by  the  following  formula  : 

(23)  Wtt  =  WLhmT  (foot-pounds) 

Wherein  W  «  =  net  work  in  foot-pounds.  W  =  weight  of 
water  or  other  liquid  pumped,  in  pounds.  LhmT  =  the  total 
measured  head  (Sec.  11),  in  feet  =  vertical  height,  in  feet, 
from  level  of  suction  supply  to  discharge  level. 

EXAMPLE.  —  A  pump  lifts  14,620  Ib.  of  water  from  a  pond  and  de- 
livers it  to  a  reservoir.  The  vertical  distance  between  the  suction-  and 
discharge-levels  is  41  ft.  What  is  the  net  pump-work?  SOLUTION.  —  By 
For.  (23),  Wu  =  WL*wr  =  14,620  X  41  =  599,420  ft.  Ib. 


32.  The  Actual  Work  Of  A  Pump  includes,  in  addition  to  the 
net  work  (Sec.  30),  all  of  the  work  performed  in  overcoming 
frictional  resistances  and  in  imparting  velocity  to  the  liquid. 


SEC.  33] 


PUMP  CALCULATIONS 


25 


Jk 

Steam  Line"-' 

Admission 
.-Port  Opens 

Admission  Port-- 
Closure 

Admission 

Exhaust 
Port 
Opens-'' 

IP 

*•'''     .-Exhaust 
/      Port 
D/      Closure 

Back- 
Pressure 
I,ne,u 

Exhaust- 
Line  

c 

Atmospheric  Line—' 

FIG.  22. — Indicator  Diagram  From  Steam-Cyl- 
inder  Of   Direct-Acting   Steam-Pump. 


The  frictional  resistances  include,  besides  the  water-friction 
in  the  suction-  and  discharge-pipes  and  in  the  pump  passages, 
the  mechanical  friction  between  the  moving  parts  of  the  pump- 
ing mechanism. 

33.  The  Rate  At  Which  A  Pump  Does  Work  May  Be 
Expressed  In  Terms  Of  Horse  Power.  —  The  total  horse 
power  developed  in  the 
steam-cylinders  of  steam- 
pumps  may  be  computed 
from  indicator  diagrams 
(Fig.  22)  taken  from  the 
steam-cylinders.  The  total 
horse  power  developed  in 
the  water-cylinders  of  re- 
ciprocating pumps  of  all 
types  may  be  computed  from  indicator  diagrams  (Fig.  23) 
taken  from  the  water-cylinders. 

EXPLANATION.  —  In  the  pump  diagram  (Fig.  23)  the  total  height, 
LhT,  indicates,  to  the  scale  of  the  diagram,  the  total  head,  For.  (6),  on 
the  pump;  this  is  called  the  indicated  head.  The  heights  Lhms  and  Lhmd 
indicate,  respectively,  the  measured  suction  head,  Lhm&  of  For.  (5),  and 
the  measured  delivery  head,  Lhmd,  For.  (5).  The  heights  s  and  d  indicate, 
respectively,  the  friction  heads  on  the  suction  and  delivery  sides  of  the 

pump.  That  is,  s  +  d  indicates 
the  total  friction  head,  Lh/T  of  For. 
(4).  The  sum  Lhms  +  Lhmd  +  d 
comprises  the  useful  head  on  the 
pump.  The  velocity  head  is 
herein  considered  as  being  so 
small  that  it  may  be  neglected. 
All  of  these  heads  are  expressed 
(Sec.  5)  in  pounds  per  square  inch. 

34.  The    Hydraulic    Or 
Water  Horse  Power  Devel- 

Fio.  23.  —  Indicator   DiagramFrom  AT*  •        u 

Water-Cylinder  Of  Reciprocating  Pump.  Oped  By  A  Pump  IS  the  US6- 
(Velocity  head  is  hereon  neglected.)  ^  horge  power  developed  in 

the  pump  cylinder  as  computed  upon  a  basis  which  comprises 
the  actual  weight  of  water  discharged  and  the  total  useful  head. 
It  may  be  expressed  by  the  following  formula: 


Pressure  Variation 
Due  to  Trembling' of 
Valves.Causeol 


Friction  Heaol  on-- 
Suet  ion  Siote 


Absolute  Zero- Pressure 


(24) 


(horsepower) 


26 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


Wherein  Puhp  =  the  theoretical  hydraulic  horse  power. 
Wim  =  the  weight  of  liquid  pumped,  in  pounds  per  minute. 
Lhu  =  the  total  useful  head,  in  feet,  =  the  head,  in  feet,  corre- 
sponding to  the  gage-pressure  (Fig.  24)  at  the  pump  discharge 
nozzle  +  the  head  due  to  the  height  of  the  discharge  nozzle 
above  the  level  of  the  source  of  suction  supply. 


ssure  forge  ;fctfe  Vctfve 
;  'Discharge 
'  Nozzle 


EXAMPLE. — A  direct-acting  steam-pump 
moves  4,160  Ib.  of  water  per  min.  against 
a  total  useful  head  of  36  ft.  What  is  the 
net  horse  power  developed?  SOLUTION. — 
By  For.  (24),  Puhp  =  W,mZW33000  =  4160 
X  36  -i-  33000  =  4.5  h.p. 

35.  The  Indicated  Efficiency  Of  A 
Reciprocating  Pump  is  the  ratio,  ex- 
pressed as  a  per  cent.,  of  the  net  useful 
horse  power  (Sec.  34)  to  the  horse 
power  computed  (Sec.  39)  from  the 
pump  indicator  diagram  (Fig.  23).  It 
may  be  expressed  by  formula: 
100PW/,P 


(25) 


Whp 


(per  cent.) 


FIG.  24. — Pump  Showing  40 
Lb.  Per  Sq.  In.  Gage  Pressure 
At  Discharge  Nozzle. 

power,    as   computed 
diagram  (Fig.  23). 


Wherein  Et-  =  the  indicated  efficiency, 
in  per  cent.  PUhP  =  the  useful  horse 
power  (Sec.  34).  PWhP  =  the  horse 
(Sec.  39)  from  the  pump  indicator 


NOTE. — The  indicated  efficiency  of  a  pump  is  a  criterion  of  the  sum 
total  of  hydraulic  losses,  or  of  the  losses  occurring  solely  in  the  water-end. 

EXAMPLE. — A  reciprocating  plunger  pump  moves  5910  Ib.  of  water 
against  a  total  useful  head  of  61  ft.  The  hydraulic  horse  power 
developed,  as  computed  from  an  indicator  diagram,  is  12.14.  What  is 
the  indicated  efficiency?  SOLUTION. — By  For.  (24),  the  useful  horse  power 
=  Puhp  =  W,mL/,u/33,000  =  5910  X  61  -=-  33,000  =  10.93  h.p.  By  For. 
(25),  the  indicated  efficiency  =  E»  =  100  PuhP/PwhP  =  100  X  10.93  -J- 
12.14  =  90  per  cent. 

36.  The  Hydraulic  Losses  Of  A  Pump  are  defined  as  those 
losses  in  hydraulic  pressure  (or  head)  which  occur  in  the 
suction  pipe  and  in  the  pump  itself.  They  comprise  pressure 
equivalents  of  the  losses  in  head  due  to:  (1)  THE  PASSAGE 


SEC.  37] 


PUMP  CALCULATIONS 


27 


OF  THE  WATER  FROM  THE  WELL  OR  OTHER  SUPPLY  SOURCE, 
THROUGH  THE  SUCTION  PIPE  AND  PUMP,  To  THE  POINT 
WHERE  THE  DISCHARGE  GAGE  (D,  Fig.  25)  Is  CONNECTED; 
these  consist  of :  (a)  suction-pipe  entrance  loss,  (b)  suction-pipe 
and  pump  velocity  loss,  (c)  suction-pipe  friction  loss,  (d)  losses 
in  suction-pipe  bends  and  connections,  (e)  friction  loss  in 
passing  through  pump  suc- 
tion valves,  (f)  friction  loss 
in  passing  through  pump 
discharge  valves.  (2)  THE 
PRESSURE  NECESSARY  To 
OVERCOME  THE  REACTION 
OF  THE  SPRINGS  OF  THE 
DISCHARGE  VALVES.  The 
pressure  lost  due  to  losses 
under  (1)  and  (2)  will  each 
be  equivalent  to  about  % 
per  cent,  of  the  total  dis- 
charge pressure,  giving  a 
total  hydraulic  loss  of 
about  1  per  cent. 


Oage  For  Indicating 
Pressures  Above  Atmos- 
pheric  Pressure '__ 


Measured  Head  Between 
Points  Of  Attachment  Of 

2S-... 


..•Steam  Connection 


_       \  Handholes 
Exhaust  \For  Access 
Connection?  To  Valves-' 

Gage  For  Indicating  Pressures  Above'' 
Or  Below  Atmospheric  Pressure 

FIG.  25. — Duplex  Fire-Pump  With  Dis- 
charge And  Suction  Gages  Attached.  De- 
signed To  Run  From  150  To  250  Ft.  Of  Piston 
Travel  Per  Min.  Steam  Cylinders,  14-In. 
Diameter.  Water  Cylinders,  8.5-In.  Diameter. 
Stroke,  12-In. 


NOTE. — In  commercial  pump 
tests  and  computations,  it  is,  as 
above  indicated,  ordinarily  un- 
derstood that  the  hydraulic 
losses  in  the  suction  pipe  are  to  be  included  with  the  losses  in  the  pump 
itself.  From  a  theoretical  standpoint,  this  is  incorrect.  But  the  pump 
manufacturers  accept  this  practice  because  it  simplifies  testing  and 
guarantees.  In  any  case,  the  true  suction-pipe  losses  are  very  small 
and  will  be  practically  the  same  for  all  pumps  which  are  doing  the  same 
work.  On  the  other  hand,  the  discharge-pipe  losses  are  never  included 
in  the  hydraulic  losses  of  a  pump. 

37.  The  Hydraulic  Efficiency  Of  A  Pump  may  be  expressed 
as  a  percentage  by  the  formula: 

100  P 

(26)  Eh  =  =T-  -=—       ,      7.    7  -  (per  cent.) 

P  -f-  the   hydraulic  losses 

Wherein:  Eh  =  the  hydraulic  efficiency,  in  per  cent. 
P  =  [pressure  as  read  on  discharge  gage  (D,  Fig.  25)  in 
Ib.  per  sq.  in.,  when  pump  is  delivering  the  quantity  of  water 
at  which  it  is  desired  to  determine  EJ  -f-  [0.433  X  (distance  in 


28 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  1 


feet  from  discharge  gage  to  surface  of  water  in  well)].  The 
hydraulic  losses  are  as  enumerated  above;  they  may  be 
obtained  as  explained  in  the  following  note. 

NOTE.  —  THE  NECESSARY  DATA  To  DETERMINE  THE  HYDRAULIC 
EFFICIENCY  of  a  given  pump  may  be  secured  in  the  following  manner: 
An  indicator  is  attached  to  the  water  cylinder  and  the  pump  driven  at 
such  a  speed  and  the  discharge  valve  is  so  throttled  that  the  pump  will 
deliver  that  quantity  of  water  at  which  the  hydraulic  efficiency  is  de- 
sired. The  discharge  valve  must  be  located  on  the  discharge  side  of 
gage  D  and  some  distance  away  from  it. 
Hydraulic  indicator  cards  (Fig.  26)  are  then 
taken,  and  at  the  instant  the  card  is  taken  the 
discharge  gage  (D,  Fig.  25)  is  read.  Compute  P 
as  above  indicated,  and  lay  off  this  pressure  (line 
AB,  Fig.  26)  to  the  scale  of  the  indicator  card, 
measuring  downward  from  the  top  of  the  indi- 

cator card  as  shown  in  Fig.  26.     The  remainder 
.. 

°*  tne  distance,  BC,  is,  to  the  scale  of  the  indi- 


-Total Hydraulic  Losses 


FIG.  26.-Indicator  Card 

Taken  On  Water  End  Of 

Steam  Pump,  Showing    cator  card,  the  hydraulic  losses,  that  is  the  pressure 

Total  Hydraulic  Losses.       required  to  overcome  the  losses. 


38.  The  Usual  Practice  In  Determining  The  Load  On  A 
Pump  is  to  attach  a  pressure-gage  to  the  discharge-pipe, 
D  (Fig.  25),  and  to  the  intake-pipe,  S  (Fig.  25),  a  gage  which 
indicates  both  vacua,  or  pressures  below  atmospheric,  and 
pressures  above  atmospheric.  Then,  if  the  pump  lifts  the 
water,  the  suction-gage,  S,  will  indicate  a  pressure  less  than 
atmospheric.  But  if  the  water  flows,  under  a  head,  to  the 
intake  of  the  pump,  the  intake-gage,  S,  will  indicate  a  pressure 
greater  than  atmospheric.  If  S  indicates  a  pressure  less  than 
atmospheric,  the  net  gage-pressure  is  found  by  adding,  to  the 
pressure  per  square  inch  shown  by  the  discharge-gage  D,  the 
pressure  per  square  inch  which  corresponds  to  the  vacuum, 
in  inches,  shown  by  the  intake  gage,  S.  If  S  indicates  a 
pressure  above  atmospheric,  the  net  gage-pressure  is  then 
found  by  subtracting,  from  the  pressure  per  square  inch 
shown  by  the  discharge-gage,  D,  the  pressure  per  square  inch 
above  atmospheric,  which  is  shown  by  the  intake-gage,  S. 
The  load  on  the  pump,  in  pounds  per  square  inch,  is  then 
equal  to  the  net  gage-pressure  plus  the  pressure  which  is  due 
to  the  hydrostatic  head,  Lhm  (Fig.  25),  between  the  points  of 
attachment  of  the  gages. 


SEC.  39]  PUMP  CALCULATIONS  29 

EXAMPLE.  —  If  the  discharge-gage,  D,  (Fig.  25)  shows  41  lb.  per  sq. 
in.,  and  the  intake-gage,  S,  shows  a  vacuum  corresponding  to  a  reduc- 
tion of  3  lb.  per  sq.  in.  below  normal  atmospheric  pressure,  then  the 
net  gage-pressure  is  41  +3  =  44  lb.  per  sq.  in.  And  if  the  vertical 
height,  Lhm,  is  5  ft.,  then  the  pressure  against  which  the  pump  works  = 
44  +  (5  X  0.433)  =  46.  1  lb.  per  sq.  in.  But  if  the  intake-gage  shows  a 
pressure  of  3  lb.  above  normal  atmospheric  pressure,  then  the  pressure 
against  which  the  pump  works  =  [41  —  3]  +  [5  X  0.433]  =  40.1  lb.  per 
sq.  in. 

NOTE.  —  CONVERSION  OP  A  VACUUM  READING  IN  INCHES  OF  MER- 
CURY To  TERMS  OF  POUNDS  PER  SQUARE  INCH  may  be  done  by  multi- 
plying the  reading  in  inches  by  0.4914,  or,  in  practice  by  0.49. 

EXAMPLE.  —  If  the-intake-gage,  5,  (Fig.  25),  shows  a  vacuum  of  5  in., 
the  difference  between  normal  atmospheric  pressure  and  the  pressure  in  the 
intake-pipe  =  5  X  0.49  =  2.25  lb.  per  sq.  in. 

39.  The  Actual  Or  Indicated  Hydraulic-  or  Water  -Horse- 
Power  Developed  By  A  Pump  is  the  total  horsepower  devel- 
oped in  the  pump  cylinder,  as  computed  upon  a  basis  of  the 
mean  pressure  throughout  the  discharge  stroke  of  the  pump 
plunger.  The  mean  pressure  is  obtained  from  an  indicator 
diagram  (Fig.  23)  taken  during  a  double  stroke  of  the  pump 
plunger.  The  hydraulic  horsepower  computed  upon  this 
basis  includes  the  power  expended  in  overcoming  all  resistance 
due  to  water-friction  from  the  inlet  orifice  of  the  suction  pipe  to 
the  outlet  orifice  of  the  discharge  pipe.  The  indicated  hydraulic 
horsepower  may  be  expressed  by  the  following  formula  : 


(27)  Pwhp  =  (H.P.);  =  ''        (horsepower) 


Wherein  Pwhp  =  the  actual  hydraulic  horsepower.  P  =  the 
load  on  the  pump,  Sec.  38,  in  pounds  per  square  inch,  which 
may  be  computed  from  the  indicator  diagram.  L/  =  the 
length  of  the  stroke,  in  feet;  A  =  the  area  of  the  plunger,  in 
square  inches.  N8  =  the  number  of  strokes  per  minute. 

40.  The  Total  Driving  Horse  Power  Developed  By  A 
Steam  Pump  Or  Delivered  To  A  Power  Pump  includes  the 
actual  hydraulic  horsepower  (Sec.  39)  plus  the  horsepower 
required  to  overcome  the  mechanical  or  metal-to-metal  friction 
in  the  complete  pumping  mechanism.  In  the  case  of  a  steam- 
pump,  the  total  driving  horse-power  will  correspond  to  the 
indicated  horse-power,  as  computed  with  the  aid  of  indicator 
diagrams  (See  the  author's  Steam  Engines)  which  is  developed 


30  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

in  the  steam  cylinder  or  cylinders.  In  the  case  of  a  power 
plunger-pump,  or  of  a  centrifugal  pump,  the  total  driving 
horsepower  will  Ibe  the  horsepower  delivered  by  belt-trans- 
mission, gear-transmission,  or  by  direct  motor-connection,  to 
the  pump  pulley,  driving-shaft  or  spindle. 

41.  The  Mechanical  Efficiency  Of  A  Reciprocating  Pump 
is  the  ratio,  expressed  as  a  per  cent.,  of  the  indicated  hydraulic 
horsepower  (Sec.  39)  to  the  driving  horsepower.  The  hydrau- 
lic horsepower  may  be  computed  from  (Fig.  23)  a  pump  indica- 
tor diagram.  The  driving  horsepower  of  a  steam-driven 
pump  may  be  computed  from  (Fig.  22)  a  steam  indicator 
diagram.  In  the  case  of  a  power  pump  (Fig.  24)  the  driving 
horsepower  is  the  total  horsepower  delivered  to  the  pump  by 
belt,  gearing,  or  direct  shaft-connection.  The  mechanical 
efficiency  may  be  expressed  by  the  following  formula  : 


(28)  Em  =  "'  =  (per  cent.) 


Wherein  Em  =  the  mechanical  efficiency,  in  per  cent.  PWhP 
=  the  hydraulic  horsepower,  as  computed  from  the  pump 
indicator  diagram.  Pbhp  =  the  driving  horsepower. 

NOTE.  —  The  mechanical  efficiency  of  a  pump  is  a  criterion  of  the  loss 
due  to  mechanical  friction  in  the  mechanism  which  transmits  the  driving 
power  to  the  water  end  of  the  pump.  The  higher  the  mechanical  effi- 
ciency the  less  the  mechanical  losses  in  the  pump. 

EXAMPLE.  —  The  hydraulic  horsepower  of  direct-acting  steam  pump,  as 
computed  from  a  pump-indicator  diagram  is  42.3.  The  driving  horse- 
power, as  computed  from  a  steam-indicator  diagram,  is  49.76.  What  is 
the  mechanical  efficiency.  SOLUTION.  —  By  For.  (28):  Em  =  100  PWhP/ 
Pbhp  =  100  X  42.3  -f-  49.76  =  85  per  cent. 

NOTE.  —  THE  MAXIMUM  MECHANICAL  EFFICIENCY  OBTAINED  WITH 
DIRECT-ACTING  STEAM  PUMPS  is  about  80  per  cent.  This  efficiency  may 
be  had  with  very  large  pumps.  The  efficiencies  diminish  with  the  sizes 
of  the  pumps.  Very  small  pumps  may  give  an  efficiency  of  only  50  per 
cent.,  or  even  less. 

42.  The  Total  Efficiency  Of  A  Pump  Is  The  Product  Of  The 
Volumetric,  Hydraulic,  And  Mechanical  Efficiencies.  It  is 

the  efficiency  which  is,  ordinarily,  specified  by  the  manufacturer 
of  the  pump.  It  is  a  criterion  of  the  pump's  overall  economy 
in  the  use  of  power.  It  may  be  expressed  by  formula  : 

/r»rk\  T-«  E^E/iEjn 

(29)  E,  =  -  (per  cent.) 


SEC.  43] 


PUMP  CALCULATIONS 


31 


Wherein  Et  =  the  total  efficiency,  in  per  cent.  Ev  =  the 
volumetric  efficiency,  in  per  cent.  Eh  =  the  hydraulic  effi- 
ciency, in  per  cent.  Em  =  the  mechanical  efficiency,  in  per 
cent. 

NOTE. — For  a  steam-driven  pump,  the  total  efficiency  recognizes  all 
losses — steam,  mechanical  and  hydraulic — from  the  steam  cylinder 
to  the  water-discharge  pipe.  For  a  power-driven  pump,  the  total  effi- 
ciency recognizes  only  the  mechanical  and  hydraulic  losses  from  the 
driven  pulley,  gear  or  shaft  to  the  water-discharge  pipe. 

43.  Total-Efficiency  Values  For  Different  Pumps  may  vary 
widely  with  the  condition  and  the  design  of  the  pump.     Cen- 
trifugal pumps  may  show  total  efficiencies  thus:-100  gal.  per 
min.,  40  per  cent.;  200  gal.,  50  per  cent.;  300  gal.,  60  per  cent.; 
400  gal.,  65  per  cent.,  600  gal.,  70  per  cent.;  800  gal.,  85  per 
cent.;  100  gal.,  75  per  cent.;  1500  gal.,  78  per  cent.     The 
efficiency  of  a  centrifugal  pump  is  also  determined  largely  by 
its  speed  and  capacity.     Hence  it  is  always  advisable,  when 
specific   data  are  required,   to  obtain  guarantees  from  the 
manufacturers.     The  total  efficiency  of  a  belt-  or  gear-driven 
power  pump  may  range  from  about  50  to  80  per  cent. 

44.  Table    Showing    Approximate    Total    Efficiencies    Of 
Steam  Pumps  In  Good  Condition  (Peele's  MINING  ENGINEERS' 
HANDBOOK). 


Total  efficiency  in  per  cent. 

Stroke 

Non- 
condensing 

Compound 
non- 
condensing 

Compound 
condensing 

Triple- 
expansion 
condensing 

4 

21 

6 

26 

26 

8 

30 

30 

10 

34 

34 

41 

50 

12 

37 

37 

45 

54 

15 

40 

40 

48 

58 

18 

43 

43 

52 

62 

24 

47 

47 

55 

66 

36 

50 

59 

70 

48 

•- 

63 

74 

32  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

45.  The  Horsepower  Required  For  Pumping  may  be  com- 
puted by  the  following  formula  : 

(30)  Pbhp  =-  ——  (horsepower) 


Wherein  Pbhp  =  the  horsepower  input  required  to  drive  a 
pump  against  the  maximum  total  head  ;  for  a  steam  pump  it  is 
the  indicated  steam  horsepower  required  for  the  steam  end, 
for  a  power-pump  it  is  the  horsepower  input  required  at  the 
driving  pulley,  gear  or  shaft.  Wim  =  the  weight  of  water  to  be 
pumped,  in  pounds  per  minute.  LHT  =  the  total  head  on  the 
pump,  in  feet.  E,  =  the  total  efficiency  of  the  pump,  in  per 
cent.,  as  defined  in  Sec  42. 

EXAMPLE.  —  It  is  required  to  pump  1,205  gal.  of  water  per  min.  against 
a  total  head  of  450  ft.  The  total  efficiency  of  the  pump  which  will  be 
used  is  64  per  cent.  What  horsepower  must  be  supplied  to  operate  the 


weigh:  1,205  X  8.3  =  10,000  lb.     By  For.  (30);  Pbhp  = 
(10,000  X  450)  -T-  (330  X  64)  =  213  h.p. 

EXAMPLE.  —  It  is  required  to  pump  10,000  lb.  of  water  per  min.  against 
a  total  head  of  450  ft.  Assuming  volumetric  and  hydraulic  efficiencies 
of  98  per  cent,  each,  and  a  mechanical  efficiency  of  80  per  cent.,  what 
horsepower  must  be  supplied? 

SOLUTION.  —  By  For.  (29),  the  total  efficiency  =  Et  =  E,-EAEm/10,000  == 
98  X  98  X  80  -T-  10,000=  77  per  cent.  By  For.  (30),  the  required  horse- 
power =  P^  =  WJmIW330  Et  =  (10,000  X  450)  -=-  (330  X  77}  =  177 
h.p. 

46.  The  Duty  Of  A  Steam  Pump  is  the  ratio  of  the  work  done 
by  the  pump  to  the  quantity  of  coal,  steam  or  heat  consumed 
in  doing  the  work. 

47.  The  Duty  Of  A  Steam  Pump  On  A  Basis  Of  Coal  Con- 
sumption may  be  found  by  the  following  formula: 


(31)  Dc  =  r         (ft    lb    per  10Q  lb    coal) 

W  c 

Wherein  Dc  =  duty,  foot  pounds,  per  100  lb.  of  coal.  W  = 
weight  of  liquid  pumped,  in  pounds.  LhT  =  total  head  on 
pump  in  feet.  Wc  =  weight  of  coal  consumed  in  pounds. 

EXAMPLE.  —  A  steam  pump  raises  12,900,000  lb.  of  water  against  a 
total  head  (Sec.  12)  of  60  feet.  The  steam  supplied  to  the  pump,  while 
doing  this  work,  requires  the  combustion  of  2,500  lb.  of  coal.  What  is  the 
duty? 


SEC.  48]  PUMP  CALCULATIONS  33 

SOLUTION.—  By  For.  (31),  Dc  =  100  WLhT/Wc  =  100  X  12,900,000  X 
60  -5-  2,500  =  30,960,000  ft.  Ib.  per  100  to.  of  coal. 

NOTE.  —  PUMP-DUTY  COMPUTED  ON  A  BASIS  OF  COAL  CONSUMPTION 
is  of  practical  use  in  comparing  the  merits  of  two  or  more  steam  pumps 
only  when  the  same  quality  of  coal  is  used  in  testing  all  of  the  pumps. 

48.  The  Duty  Of  A  Steam  Pump  On  A  Basis  Of  Steam 
Consumption  may  be  computed  by  the  following  formula  : 


(32)  Ds  =  (ft>  lb>  per  1000  lb  steam) 

ws 

Wherein  Ds  =  duty,  in  foot  pounds  per  1,000  lb.  of  dry  steam. 
W  =  weight  of  water  pumped,  in  pounds.  LhT  =  total 
head  on  pump  in  feet.  Ws  =  weight  of  steam  consumed, 
in  pounds. 

EXAMPLE.  —  A  steam-pump  raises  8,765,000  lb.  of  water  against  a 
total  head  of  125  ft.  The  steam-consumption  is  8,315  lb.  What  is  the 
duty? 

SOLUTION.—  By  For.  (32)  Ds  =  1,000  WLhT/Ws  =  1,000  X  8,765,000  X 
125  -5-  8,315  =  131,764,883  ft.  lb.  per  1,000  lb.  of  dry  steam. 

NOTE.  —  PUMP-DUTY  COMPUTED  ON  A  BASIS  OF  STEAM-CONSUMPTION 
may  have  only  an  approximate  value.  This  may  be  due  to  the  difficulty 
of  determining  the  exact  weight  of  dry  steam  used.  It  may  also  be  due  to 
variations  of  steam  pressure.  A  given  weight  of  high-pressure  steam 
will  do  more  work  in  the  cylinder  than  the  same  weight  of  compara- 
tively low  pressure  steam. 

49.  The  Duty  Of  A  Steam  Pump  On  A  Basis  Of  The  Quan- 
tity Of  Heat  Consumed  may  be  computed  by  the  following 
formula  : 

,QQv     p.       l,000,000(P<f  +  Pi  +  Pp)ALfN8       1,  000,000  WL^r 

(33)  Dh  =  H  ~^j~ 

(ft.  lb.  per  1,000,000  B.t.u.) 

Wherein  DA  =  duty,  in  foot  pounds,  per  1,000,000  B.t.u. 
Pd  =  discharge  pressure,  in  pounds  per  square  inch,  as  indicated 
by  a  gage  in  the  discharge  pipe.  Pi  =  intake  pressure,  in 
pounds  per  square  inch,  as  measured  from  atmospheric  pres- 
sure (Sec.  38)  by  a  gage  in  the  intake  pipe  —  to  be  added  if 
negative  and  to  be  subtracted  if  positive.  PD  =  pressure,  in 
pounds  per  square  inch,  due  to  hydrostatic  head  between 
points  of  attachment  of  pressure  gages.  A  =  effective  area  of 
plunger,  in  square  inches.  L/  =  length  of  stroke,  in  feet.  N9 


34  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

=  total  number  of  strokes  H  =  total  quantity  of  heat 
consumed,  in  British  thermal  units,  as  determined  by  steam 
consumption  test;  see  the  author's  STEAM  ENGINES. 

NOTE. — PUMP-DUTY  COMPUTED  ON  A  BASIS  OF  HEAT-CONSUMPTION 
is  more  nearly  exact  than  computations  (Sec.  49)  on  bases  of  coal-  or 
steam-consumption.  Since  the  determining  factor  is  the  actual  quan- 
tity of  heat  energy  expended  in  the  steam-cylinder,  pump-duty  figured 
on  this  basis  provides  a  true  criterion  of  the  comparative  working  effi- 
ciencies of  two  or  more  different  pumps.  This  method  has  been  recom- 
mended by  the  A.  S.  M.  E. 

EXAMPLE. — A  duplex  steam-pump  has  inside-packed  plungers  of  20- 
in.  diameter  and  15-in.  stroke.  The  plunger-rods  are  of  3-in.  diameter. 
The  total  heat  in  the  steam  supplied  to  this  pump,  during  a  duty  trial, 
was  17,642,400  B.t.u.  The  pump  made,  during  the  trial,  37,264  strokes. 
The  average  discharge-pressure,  as  indicated  by  a  gage  in  the  discharge 
pipe,  was  96  Ib.  per  sq.  in.  The  average  intake-pressure,  as  indicated 
by  a  gage  in  the  suction  pipe,  was  4  Ib.  per  sq.  in.  below  atmospheric 
pressure.  The  pressure  due  to  the  hydrostatic  head  between  the  suction- 
and  discharge-gages  was  3.5  Ib.  per  sq.  in.  What  was  the  duty? 

SOLUTION.— By  For.  (33), DA  =  1,000,000  (Pd  ±  P.  +  PD)  ALfNt/H  = 
1,000,000  X  (96  +  4  +  3.5)  X  [202  X  0.7854  -  (32  X  0.7854  -^  2)]  X 
(15  -J-  12)  X  37,264  -r-  17,642,400  =  84,884,000  ft.  Ib.  per  1,000,000 
B.t.u. 

50.  The  Miscellaneous  Reciprocating-Pump  Formulas 
which  follow  supplement  those  given  previously  herein. 
These  formulas  relate  specifically  to  single-acting  simplex 
pumps.  The  number  of  strokes  per  minute  =  the  number  of 
pumping  strokes  per  minute  =  %  the  number  of  reversals  of 
the  piston.  Where  cylinder  area  is  used  in  the  following 
formulas,  it  means  the  cross-sectional  area  of  the  cylinder 
taken  at  right  angles  to  the  piston  rod. 

NOTE. — IN  THE  EVENT  THAT  THESE  FORMULAS  ARE  USED  IN  DOUBLE- 
ACTING-PUMP  COMPUTATIONS,  the  number  of  working  strokes  per  minute  = 
the  number  of  reversals  per  minute  of  the  piston.  Also,  in  double-acting- 
pump  computations,  for  cylinder  area  must  be  substituted  [cylinder  area  — 
(piston-rod  area  -r-  2)1.  For  (diameter  of  cylinder)2  must  be  substituted 
{ (diameter  of  cylinder)2  —  [(diameter  of  piston  rod)2  -f-  2]  [ . 

(34)  Gal.  per  min. 

_  (Strokes  per  min.)  X  (Stroke  in  in.)  (Dia.  of  water  cyl.  in  in.)  2 

294 


SEC.  50]  PUMP  CALCULATIONS  35 

EXAMPLE. — How  many  gallons  of  water  will  be  delivered  per  minute 
by  a  pump  having  a  water  cylinder  8  in.  in  diameter  by  12  in.  stroke 
when  it  is  making  100  strokes  per  minute?  SOLUTION, — Gallons  per 
minute  =  (100  X  12  X  8  X  8)  -3-  294  =  261  gal  per  min. 

(35)  Dia.  of  water  cylinder  in  in. 


Gal.  per  min 


(Stroke  inin.)  X  (Strokes  per  min.) 

EXAMPLE. — What  will  be  the  required  water-cylinder  diameter  to 
pump  200  gal.  per  min.,  if  the  length  of  stroke  is  10  in.  and  the  pump 
makes  120  strokes  per  min.?  SOLUTION. — Diameter  of  water  cylinder 
in  inches  =  17. 14  \/ (200)  -J-  (10  X  120)  =  7  in. 

(36)  Area  of  water  cylinder  in  sq.  in. 

=  (231)  X  (Gal  per  min.) 

(Strokes  per  min.)  X  Stroke  in  in.) 

EXAMPLE. — What  area  of  water  cylinder  is  required  to  pump  330  gal. 
per  min.,  if  the  pump  has  a  16  in.  stroke  and  makes  80  strokes  per  min.? 
SOLUTION.— Area  of  water  cylinder  =  (231  X  330)  -s-  (80  X  16)  =  59.6 
sq.  in. 

(37)  Area  of  water  cylinder  in  sq.  in. 

(3.85)  X  (Gal,  per  hr.) 

(Strokes  per  min.)  X  (Stroke  in  in.) 

EXAMPLE. — A  pump  has  a  stroke  of  24  in.,  and  makes  50  strokes  per 
min.  What  must  be  the  water-cylinder  area  if  it  is  to  pump  97,920  gal. 
per  hr.?  SOLUTION. — Area  of  water  cylinder  =  (3.85  X  97,920)  -5- 
(50  X  24)  =  314  sq.  in. 

(38)  Length  of  stroke  in  in. 

(231)  X  (Gal,  per  min.) 

(Strokes  per  min.)  X  (Area  of  water  cyl.  in  sq.  in.) 

EXAMPLE. — What  must  be  the  length  of  stroke  of  a  pump  having  a 
water-cylinder  area  of  28.3  sq.  in.,  if  it  must  pump  146  gals,  per  min.  when 
making  120  strokes  per  minute?  SOLUTION. — Length  of  stroke  =  (231  X 
146)  +  (120  X  28.3)  =  10  in. 

(39)  Stroke  in  in. 

=  (Gal,  per  hr.)  X  (4.9) 

(Strokes  per  min.)  X  (Diam.  of  water  cylinder  in  in.)2 

EXAMPLE. — What  will  be  the  required  length  of  stroke  to  pump  35,251 
gal.  per  hr.  if  the  pump  has  a  water  cylinder  12  in.  in  diameter  and  makes 
80  strokes  per  min.?  SOLUTION.— Length  of  stroke  =  (35.251  X  4.9)  -r 
(80  X  12  X  12)  =  15  in. 


36  STEAM  POWER  PLANT  AUXILIARIES  [Div.  I 

(40)  Stroke  in  in. 

=  (Gal,  per  min.)  X  (294) 

(Strokes  per  min.)  X  (Diam.  of  water  cyl.  in  in.)2 

EXAMPLE. — What  will  be  the  required  length  of  stroke  to  pump  587  gal. 
per  min.  if  the  pump  has  a  water  cylinder  12  in.  in  diameter  and  makes 
66.6  strokes  per  min.  SOLUTION. — Length  of  stroke  =  (587  X  294)  -r- 
(66.6  X  12  X  12)  =  18  in. 

//11N  c,    1  .  (Gal.  perhr.)  X  (3.85) 

(4 1 )  Strokes  per  mm.  =  ,„,  .      ^-7—    — : —  '  .    .v        •  — ; — : — — . 

(Water-cyl.  area  in  sq.  in.)  X  (Stroke  in  in.) 

EXAMPLE. — How  many  strokes  per  minute  will  a  pump  have  to  make 
to  pump  8,812  gal.  per  hr.  if  it  has  a  water-cylinder  area  of  28.3  sq.  in. 
and  a  length  of  stroke  of  12  in.?  SOLUTION. — Number  of  strokes  = 
(8,812  X  3.85)  -:-  (28.3  X  12)  =  100  per  min. 

(Gal  per  hr.)  X  (4.9) 

(42)  Strokes  per  mm.  =  -^- — -, — .    .    .  ^ ,  ^ .    '  —r-. — ^-^ 

(Stroke  in  in.)  X  (Dia.  of  water  cyl.  in  in.) 2 

EXAMPLE. — How  many  strokes  per  minute  will  a  pump  have  to  make 
to  pump  8,812  gal.  per  hr.  if  it  has  a  water-cylinder  diameter  of  6  in.  and 
a  length  of  stroke  of  12  in.?  SOLUTION. — Number  of  strokes  =  (8,812  X 
4.9)  -f-  (12  X  6  X  6)  =100  per  min. 

(43)  Strokes  per  min. 

_  (Gal,  per  min.)  X  (2  3  1) 

(Stroke  in  in.)  X  (Area  of  water  cyl.  in  sq.  in.) 

EXAMPLE. — How  many  strokes  must  a  pump  make  per  minute  to 
pump  146  gal.  per  min.  if  it  has  a  water-cylinder  area  of  28.3  sq.  in.  and 
a  10  in.  stroke?  SOLUTION. — Number  of  strokes  =  (146  X  231)  -5- 
(10  X  28.3)  =  120  per  min. 

(44)  Water-gage    pressure    necessary    to    balance    steam-gage 

(Steam-gage  pressure) (Diam.  in  in.  of  steam-cyl.)2 

pressure  =  -  — 77^ —  — - — : 7 : rr^ — 

(Diam.  in  in.  of  water-cyl.)2 

EXAMPLE. — If  a  pump  has  a  steam  cylinder  5  in.  in  diameter  and  a 
water  cylinder  3  in.  in  diameter,  what  water-gage  pressure  will  be  re- 
quired to  balance  a  steam-gage  pressure  150  Ibs.  per  sq.  in.?  SOLUTION. 
— Water-gage  pressure  =  (150  X  5  X  5)  -i-  (3  X  3)  =  416  Ibs.  per  sq.  in. 

(45)  Steam-gage     pressure    necessary    to     balance    water-gage 

(Water-gage  pressure) (Dia.  in  in.  of  water  cyl.)2 

pressure  =  —  — 7^^ — : — : 7~; —      — 7T5 

(Dia.  in  in.  of  steam  cyl.)2 

EXAMPLE. — If  the  water  cylinder  of  a  pump  is  8  in.  in  diameter  and 
the  steam  cylinder  is  12  in.  in  diameter,  what  must  be  the  steam-gage 
pressure  in  order  to  just  balance  a  water-gage  pressure  of  130  Ibs.  per 
sq.  in.?  SOLUTION.— Steam-gage  pressure  =  (130  X  8  X  8)  -f-  (12  X  12) 
=  57.8  Ibs.  per  sq.  in. 


SEC.  50]  PUMP  CALCULATIONS  37 

(46)  Area  of  water  cylinder  in  sq.  in.  necessary  to  balance  a 
given  steam  pressure  = 

(Area  of  steam  cyl.  in  sq.  in.)  X  (steam  pressure  in  Ibs.  per  sq.  in.) 
(Water  pressure  in  Ibs.  per  sq.  in.) 

EXAMPLE. — A  pump  has  a  steam-cylinder  area  of  113.1  sq.  in.  If  the 
steam  gage  reads  60  Ibs.  per  sq.  in.  and  the  water-pressure  gage  reads 
135  Ibs.  per  sq.  in.,  what  must  be  the  area  of  the  water  cylinder  if  the 
piston  is  just  balanced?  SOLUTION. — Area  of  water  cylinder  =  (113.1  X 
60)  -r  135  =  50.25  sq.  in. 

(47)  Area  of  steam  cylinder  in  sq.  in.  necessary  to  balance  a 
given  water  pressure  = 

(Area  of  water  cyl.  in  sq.  in.)  X  (Water  pressure  in  Ibs.  per  sq.  in.) 
(Steam  pressure  in  Ibs.  per  sq.  in.) 

EXAMPLE. — A  pump  has  a  water-cylinder  area  of  50.25  sq.  in.  If  the 
water  gage  shows  a  pressure  of  135  Ibs.  per  sq.  in.  and  the  steam  gage 
shows  a  pressure  of  60  Ibs.  per  sq.  in.,  what  must  be  the  area  of  the 
steam  cylinder  if  the  piston  is  just  balanced?  SOLUTION. — Area  of 
steam  cylinder  =  (50.25  X  135)  -f-  (60)  =  113.1  sq.  in. 

QUESTIONS  ON  DIVISION  1 

1.  What  conditions  govern  the  height  to  which  water  may  be  lifted  by  pump-suction? 
What  is  the  practical  limit  of  suction-lift  at  sea-level?     What  is  the  practical  limit  of 
temperature  at  which  water  may  be  lifted  by  pump-suction? 

2.  What  is  a  static  head?     What  is  its  significance? 

3.  Why  should  water  from  an  open  heater  enter  the  suction-nozzle  of  a  boiler  feed 
pump  under  a  static  head?     Describe  the  action  that  may  occur  within    the    pump 
cylinder,  if  the  inlet  static-head  is  insufficient. 

4.  Enumerate  the  three  general  forms  of  resistance,  or  head,  which  must  be  overcome 
in  pump-operation.     Which  of  these  comprise  the  dynamic  head? 

5.  What  is  velocity-head?     Friction-head?     Measured-head? 

6.  Enumerate  the  causes  of  friction-head. 

7.  What  is  entrance-head? 

8.  If  a  pump  is  discharging  into  the  compression-tank  of  an  elevator  system,  how  is 
the  head  due  to  the  gage-pressure  in  the  tank  classified  in  computations  relating  to  the 
performance  of  the  pump? 

9.  What  is  the  total  head  on  a  pump? 

10.  Do  computations  based  upon  values  taken  from  published  tables  afford,  in  all 
cases,  accurate  criteria  of  the  water-friction  in  pipes?     Why? 

11.  What  is  the  displacement  of  a  reciprocating  pump? 

12.  What    constitutes    the    effective   displacement   area    of    an    outside-end-packed 
plunger?     Of  a  center-packed  plunger?     Of  an  inside-packed  plunger  or  piston? 

13.  What  is  pump-slip?     Under  what  circumstances  may  pump-slip  be  negative? 

14.  Explain  the  influence  of  high  piston  speed  on  pump-slip. 

15.  What  is  meant  by  the  volumetric  efficiency  of  a  pump? 

16.  What  should  be  the  maximum  limit  of  piston-speed  for  a  pump  vrith  a  20-in. 
stroke?     With  a  9-in.  stroke?     With  a  3-in.  stroke? 

17.  What  is  meant  by  the  useful  work  of  a  pump?     The  actual  work? 

18.  What  is  meant  by  the  indicated  efficiency  of  a  reciprocating  pump?     What  losses 
does  this  efficiency  particularly  signify? 


38  STEAM  POWER  PLANT  AUXILIARIES  [Div.  1 

19.  What  is  meant  by  the  hydraulic  efficiency  c/«  pump? 

20.  What  constitutes  the  total  head  in  determining  the  hydraulic  efficiency? 

21.  Describe  an  experimental  method  of  determining  the  load  on  a  pump. 

22.  What  is  meant  by  the  mechanical  efficiency  of  a  reciprocating  pump?     What  loss 
is  determined  by  this  efficiency? 

23.  What  is  meant  by  the  total  efficiency  of  a  pump?     What  does  this  efficiency 
signify? 

24.  What  is  meant  by  the  duty  of  a  steam  pump? 

25.  What  conditions  may  vitiate  the  practical  significance  of  pump-duty  computed 
on  a  basis  of  coal-consumption?     On  a  basis  of  steam-consumption? 

26.  Wherein  lies  the  practical  value  of  pump-duty  computed  on  a  basis  of  heat- 
consumption? 

PROBLEMS  ON  DIVISION  1 

1.  Atmospheric  pressure  at  an  altitude  of  13,000  ft.  above  sea-level  is  approximately 
9  Ib.  per  sq.  in.     What  is  the  practical  suction  lift  at  this  elevation? 

2.  A  direct  acting  steam  pump  is  lifting  water  through  a  height  of  11  ft.  and  dis- 
charging it  through  an  additional  height  of  19  ft.     What  is  the  total  static  head,  ex- 
pressed in  terms  of  pressure? 

3.  A  boiler  feed  pump  has  the  water  fed  to  it  (Fig.  3)  by  gravity.     It  is  assumed  that 
the  inlet  head  thus  produced  is  wholly  expended  in  filling  the  pump  cylinder  with  water 
against  a  tendency  of  the  water,  due  to  its  temperature,  to  vaporize  in  the  cylinder. 
Hence  no  part  of  this  head  is  available  for  balancing  an  equivalent  head  on  the  delivery 
side.     The  delivery  pipe  is  of  !J-£  in.  size.     It  has  a  total  horizontal  length  of  115  ft. 
and  a  vertical  length  of  38  ft.     There  are  three  90  deg.  elbows,  two  plugged  tees  and 
two  globe  valves  in  the  line.     The  boiler  pressure  is  150  Ib.  per  sq.  in.     If  about  20  gal. 
of  water  are  delivered  per  minute,  what  pressure  head  will  be  necessary  in  the  pump 
cylinder?     What  is  the  equivalent  gage  pressure? 

4.  If  all  conditions  remain  the  same  as  in  prob.  3  except  that  the  pipe-size  is  changed 
to  1  in.,  how  many  gallons  of  water  will  be  delivered? 

5.  A  direct-acting  simplex  steam  pump  is  required  to  deliver  90  cu.  ft.  of  water  per 
miiH     The  flow  velocity  in  the  suction  pipe  is  assumed  to  be  210  ft.  per  min.  and  in  the 
discharge  pipe  390  ft.  per  min.     What  should  be  the  sizes  of  the  piping  for  suction  and 
discharge? 

6.  An  outside  end-packed  duplex  plunger  pump  has  plungers  of  10  in.  diameter. 
The  stroke  is  20  in.     Each  plunger  makes  65  strokes  per  min.     What,  if  the  pump  is 
double  acting,  is  the  displacement  in  cubic  feet  per  minute? 

7.  The  displacement  of  a  pump  is  510  cu.  ft.  per  min.     The  pump  delivers  487  cu. 
ft.  of  water  per  min.     What  is  the  slip? 

8.  What  is  the  volumetric  efficiency  of  the  pump  of  Prob.  7? 

9.  The  plunger  diameter  in  a  direct  acting  simplex  steam  pump  is  3.5  in.     The  stroke 
is  6.5  in.     When  the  plunger  makes  110  strokes  per  minute,  the  volumetric  efficiency  is 
98  per  cent.     What  is  the  discharge? 

10.  A  direct  acting  duplex  steam  pump  is  required  to  deliver  990  cu.  ft.  of  water  per 
hr.  while  running  100  ft.  of  piston  travel  per  min.     If  the  volumetric  efficiency  is  96 
per  cent,  what  should  be  the  water-piston  diameter? 

11.  The  water  piston  diameter  in  a  direct-acting  steam  pump  is  5  in.     The  pump 
discharges  through  a  2  in.  pipe.     The  piston  travel  is  80  ft.  per  min.     What  is  the 
velocity  of  flow  in  the  discharge  pipe? 

12.  A  pump  elevates  20,106  Ib.  of  water  per  minute  through  a  total  vertical  height 
of  38.5  ft.     What  net  work,  in  foot  pounds,  is  done  in  one  minute? 

13.  What  is  the  net  horse  power  expended  by  the  pump  in  Prob.  12? 

14.  What  is  the  horse  power  required  for  lifting  9,500  Ib.  of  water  per  minute  against 
a  useful  head  of  310  ft.  when  the  total  efficiency  of  the  pump  is  85  per  cent.? 

15.  A  steam  pump  elevates  9,000,000  Ib.  of  water  against  a  total  useful  head  of  120  ft. 
The  coal  consumption  of  the  boilers  while  furnishing  steam  for  this  work  is  3,500  Ib. 
What  is  the  duty  of  the  pump  per  100  Ib.  of  coal? 


DIVISION  2 
DIRECT-ACTING  STEAM  PUMPS 

51.  Direct-Acting  Steam  Pumps  For  Modern  Power-Plant 
Service  are  (Fig.  27)  of  the  reciprocating  double-acting, 
suction  type.  That  is,  they  are  designed  to  raise  water  by 
suction  from  a  lower  level,  and  to  deliver  it  during  each  stroke 
of  the  moving  element  (Fig.  28)  to  tanks,  boilers,  or  wherever 
else  required. 


Metal  Snap- 
P/'ngrs-^ 


Brass  Liner  Forced  < 

into  Cylinder  Under  ; 

Pressure '* 


Fia.  27. — Water-End  Of  Direct-Acting  Steam- Pump  Having  Water-Piston  Fitted  With 

Snap  Rings. 

EXPLANATION. — The  movement  toward  the  left  of  the  piston  (P-Fig. 
28)  as  indicated,  causes  the  water  in  the  space  B  to  be  forced  out  through 
the  left-hand  pair  of  discharge  valves,  VD.  Coincidentally,  it  creates  a 
partial  vacuum  in  the  space  A.  That  is,  it  causes  the  air  pressure  in 
the  space  A  to  be  lowered.  This  reduction  of  pressure,  per  square  inch, 
must  be  equal  to,  or  greater  than,  the  pressure  per  square  inch  which  is 
imposed  by  the  weight  of  a  column  of  water  of  the  height  LhS  of  Fig.  1. 
The  external  atmospheric  pressure  will  then  force  the  water  up  the  suc- 
tion pipe,  S,  and  through  the  right-hand  pair  of  suction  valves,  Vs.  On 
the  return  stroke,  a  partial  vacuum  is  created  in  the  space  B.  Water 
then  enters  space  B  through  the  left-hand  pair  of  suction  valves,  while 
the  water  in  A  is  forced  out  through  the  right-hand  pair  of  discharge 
valves. 

39 


40 


STEAM  POWER  PLANT  AUXILIARIES 


[Dry.  2 


NOTE. — The  intake  water  often  flows  under  pressure  to  the  suction 
nozzles  (Sec.  4)  of  power-plant  pumps.  The  intake  pressure  may  be 
due  (Fig.  29)  to  an  elevated  source  of  supply,  or  it  may  be  derived  from 
street  mains. 


Discharge  Prpe 


Float  for- 
Operating 

Steorm  Vafve-\-  - ' 


Sucthn  Pipe-  - 


Fid.     28.— Illustrating   The   Principle   Of 
The  Reciprocating  Pump. 


.........  , 

Sucthn  from  Wei/:1 


FIG.  29.  —  Intake  Water  Often  Flows  Un- 
der Pressure  To  The  Suction  Nozzle. 


62.  The  Allowable  Velocity  Of  Flow  In  The  Water-Piping 
Of  A  Direct  -Acting  Steam -Pump  is:  (1)  For  the  intake-pipe, 


Packing..^ 


Discharge 
Nozzle-.^ 


"  ••  $LKf ion  Nozzle  :.-'•-••- 
Fia.  30. — Outside  Center-Packed  Plunger-Pump. 

about  200  ft.  per  min.  (2)  For  the  discharge-pipe  of  a  single 
pump,  about  400  ft.  per  min.  (3)  For  the  discharge-pipe  of  a 
duplex  pump,  about  500  ft.  per  min.  (4)  For  the  centrifugal 


SEC.  53] 


DIRECT-ACTING  STEAM  PUMPS 


41 


pump  about  600  ft.  per  min.  in  both  the  discharge  and  suction 
pipes. 


.-Plunger 


.•Cylinder 


Packing--' 


FIG.   31. — An  Outside  End-Packed  Plunger-Pump. 

53.  Direct-Acting  Steam  Pumps  May  Be  Classified,  With 
Reference  To  Their  Water -Ends,  as  follows:  (1)  PISTON- 
PUMPS  (Fig.  27).  (2)  PLUNGER-PUMPS  (Fig.  30).  The  latter 


Eye-Bolt 


Discharge 
Outlet*. 


Discharge, 

Nozz'e-- 

Dischargz 

Vctlvus--- 


Fia.     32. — Pump-Plunger    Inside-Packed 
With  Metal  Ring. 


'-Suction  Inlet 


FIG.  33.— Water-End  Of  Direct-Act- 
ing Steam-Pump  With  Fibrous-Packed 
Water-Piston. 


may  be  subdivided  into :  (a)  Outside  end-packed  plunger-pumps 
(Fig.  31).  (6)  Outside  center-packed  plunger-pumps  (Fig.  30). 
(c)  Inside-packed  plunger-pumps  (Fig.  32).  In  a  piston-pump, 
the  piston  traverses  a  liner  or  barrel  (Fig.  33)  which  is  com- 


42 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


monly  made  of  brass.  The  liner  may  be  secured  (Fig.  27)  by 
means  of  a  force-fit  with  the  bore  of  the  iron  cylinder  casting, 
or  (Fig.  33)  by  means  of  stud-bolts  or  cap-screws.  A  tight 
joint  between  the  periphery  of  the  piston  and  the  bore  of  the 
liner  is  obtained  (Fig.  33)  by  means  of  rings  of  square  fibrous 
packing,  or  (Fig.  27)  by  using  metal  snap-rings.  In  a  plunger- 
pump  either  end-packed,  center-packed  or  inside-packed,  the 
plungers  pass  through  fibrous — or  metal-packed  stuffing-boxes. 

NOTE. — PISTON-PUMPS  MAY  BE  USED  AGAINST  DISCHARGE-HEADS 
UP  To  ABOUT  300  LB.  PER  SQ.  IN.  (Sec.  38).  Difficulty  may  be  had, 
however,  in  keeping  the  piston  tightly  packed  if  the  head-pressure  ex- 
ceeds 150  Ib.  per  sq.  in.  The  fact  that  the  packing  is  stationary  in  the 


PumpCylirKkr.-Pump  cylinder 


•Piston  Rfnys  "Binding NUT   '-Water  Grooves 

Fia.  34. — Metal-Packed  Pump-Piston.        FIG.  35. — Water-Packed  Pump-Piston. 

plunger  pumps  and  that  it  may  be  tightened  more  readily,  renders  it 
more  effective  therein  than  in  piston  pumps.  For  low  pressures,  the 
piston  are  less  expensive  than  the  plunger  pumps. 

PLUNGER-PUMPS  ARE  COMMONLY  USED  AGAINST  DISCHARGE-HEADS 
ABOVE  ABOUT  200  LB.  PER  SQ.  IN.  For  pressures  above  300  Ib.  per  sq. 
in.,  choice  of  plunger-pumps  as  against  piston  pumps,  is  practically 
imperative. 

WATER-PISTONS  PACKED  WITH  METALLIC  RINGS  (Fig.  34)  are  com- 
monly used  in  hot-water  pumps.  Water-packed  pistons  (Fig.  35)  are 
also  sometimes  used  for  hot-water  service.  The  packing  is  afforded  by 
the  water  which  becomes  pocketed  in  a  series  of  annular  grooves  in  the 
piston's  periphery. 

54.  The  Water-Piston  Packing  In  Direct-Acting  Steam 
Pumps  For  Power -Plant  Service  is  generally  fibrous.  It  is 
commonly  known  as  canvas  or  duck  hydraulic-packing.  It 
consists  mainly  of  cotton  fiber  (Fig.  36)  interlaid  with  a 
rubber  composition.  Its  cross-section  is  square. 


SEC.  55] 


DIRECT-ACTING  STEAM  PUMPS 


43 


NOTE. — RINGS  OF  CANVAS  PACKING  FOR  A  PUMP-PISTON  should 
(Fig.  36)  be  cut  about  2{6  in-  snort  of  meeting  when  inserted  (Fig.  37) 
in  the  cylinder-bore.  Also,  the  joints  should  be  lapped  (Fig.  36).  This 
packing  is  commonly  made  in  layers.  The  layers  can  (Fig.  38)  be  peeled 
off  to  get  rings  of  suitable  width  or  thickness.  If  the  packing  is  too  deep 


.Flcmgect Enof  -Bronze     r,,/,vw,rr/vcf;rw  Pine 
of  Piston    I  Liner  yCylmofer  Casting  stick 


Fia.  36.— Ring  Of  Tuck  Canvas 
Piston-Packing. 


Packing  Recess 


Fia.  37. — How  Packing   Is  Inserted  In  Packing 
Recess  Of  Pump-Piston. 


to  fit  the  recess  around  the  piston,  it  may  be  cut  down.  A  convenient 
and  accurate  method  (Fig.  39)  of  doing  this  is  by  gripping  the  packing 
in  a  vise  and  paring  it  with  a  sharp  drawing  knife.  The  rings  should  be 
well  coated  with  graphite  and  cylinder  oil.  They  should  be  just  tight 
enough  to  require  moderate  pressure  of  the  fingers  to  force  them  into 
the  recess  around  the  piston.  They  may  then  be  forced  home  (Fig. 


'md^  tfrrjtf  Peeled  Down 


Canvas  Packing 


Ktnys  tfTurtis  ''follower 
Packing....'      Plate 


FIG.  38. — A  Canvas-Packed  Pump        FIG.  39. — Bow  Depth  Of  Canvas  Piston-Packing 
Piston.  May  Be  Cut  Down. 

37)  with  a  stick  of  soft  wood.  Rings  of  canvas  packing  may  be  partially 
expanded  to  their  working  size,  before  inserting  them  in  the  piston-recess, 
by  soaking  them  for  a  few  hours  in  warm  water. 

55.  The  Valves  Of  Power -Plant  Pumps  are  generally  of  the 
poppet-disc  type  (Figs.  40,  41,  42,  and  43),  rising  vertically 
from  flat  seats.  Conical-seated  valves  (Fig.  44)  are  also  used. 


44 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


The  discs  (D-Fig.  40)  of  flat-seated  valves  are  commonly  made 
of  a  composition  of  rubber  with  certain  other  substances. 
They  are  also  made  of  metal  (Fig.  43).  Usually  a  brass  cap 


Lock-Nut- 

Brass  Cop-, 


Spring-Nut 
.-Spring 
$  /  .-Valve  Seat 
Cast  I 


Stud- 


Brass     Valve 
Ptorre*    Deck- 


FIG.  40.  —  Flat-Seated  Pump-  Valve  With        FIG.  41.  —  Rubber  Pump-Valve  Flat-Disk 


Composition  Rubber  Disc. 


Type. 


Brass- 
Valve  \ 
Disc  \ 


Screw 
Threads 


Valve-. 
Deck  \ 

•Stud    • 


Bushing 


Fia.    42. — Bronze  Pump- Valve  Flat- 
Disk  Type. 


FIG.  43. — Sectional  Elevation  Of  Bronze 
Pump-Valve. 


.•Spring  Follower., -jock- Nut 
' 


Brass  Sea 

Bushing-  •  •  Vfcy/ve  0ecA; 

FIG.  44.— Conical-Seated  Pump- Valve  With  Brass  Disc. 

or  plate  (P-Fig.  40)  is  used  to  stiffen  the  rubber  disc  and 
prevent  .warping.  It  also  serves  to  protect  the  disc  from 
the  direct  thrust  of  the  spring  ($-Fig.  40).  The  discs  of 
conical-seated  valves  (Fig.  44)  are  generally  made  of  metal. 


SEC.  55] 


DIRECT-ACTING  STEAM  PUMPS 


45 


NOTE. — THE  HARDNESS  OF  RUBBER  COMPOSITION  VALVE-DISCS  should 
be  adapted  to  the  special  requirements  of  the  service  for  which  the  discs 
are  intended.  The  valve-discs  of  vacuum  pumps  should  be  soft  and 
pliable.  Such  discs  are  also  suitable  for  pumps  working  against  water 
pressures  up  to  about  75  Ib.  per  sq.  in.  For  pressures  from  about  75 
to  150  Ib.  per  sq.  in.,  hard  rubber  composition  discs  usually  give  the  best 


Winy-  Type  Discharge  Valves-.,. 


-Outlet 

_..-- Diaphragm 


\\\  \\  \  \  \ \\  \  \ .\.\\\\  \ \\  \  \  \  \  \\  x \  \  \  \  \  \  \  \\ 

FIG.  45. — Water-End  Of  Direct-Acting  Steam-Pump  For  Hydraulic-Pressure  Service. 


service.  For  pressures  from  about  150  to  300  Ib.  per  sq.  in.,  specially- 
hard  vulcanized  rubber  composition  valve-discs  generally  suffice.  Metal 
valve-discs  are  required  for  pressures  above  about  300  Ib.  per  sq.  in. 
The  hardness  of  valve-discs  should  also  depend  on  the  temperature  of 
the  water  pumped.  The  higher  the  temperature  of  the  pumped  water, 


Packing  Gland- 


\\\\  \W'> 

-    '  N      '  •  '  '  % 


Seat  Bus  hing-  j- > 


FIG.  46.—  Water-End  Of  Duplex  Outside-        FIG.  47.—  Ball   Pump-  Valve   For  High 


Packed     Plunger-Pump     Equipped      With 
Pot  Valves. 


Pressure  Service. 


the  harder  the  valve  discs  should  be.  Metal  valve  discs  are  frequently 
used  for  hot-water  service. 

THE  SEATS  OF  METAL-DISC  PUMP-VALVES  SHOULD  BE.  OF  THE 
SAME  KIND  OF  METAL  As  THE  Discs.  This  is  to  prevent  electrolytic 
action. 

Wing-valves  are  commonly  used  in  high-pressure  pumps  (Figs.  45  and 


46 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


46)  of  the  pot-valve  type.  Ball-valves  (Fig.  47)  are  also  sometimes  used 
in  pumps  of  this  type.  A  valve  which  is  commonly  used  in  pumping 
clear  liquids  is  shown  in  Fig.  48. 


Composition 
.-Valve  Disc 


•Stem 

..Conical  Spring 
Sleere 


Seat 
•Bushing 


Stud* 


FIG.    48. — Type    Of    Pump    Valve 
Used  For  Clear  Liquids. 


FIG.  49. — Kinghorn  Pump  Valve. 


56.  The  Kinghorn  Valve  For  Air  Pump  Service  (Fig.  49) 
consists  of  three  bronze  disks,  each  about  J-^2  in.  thick,  which 
are  mounted  loosely  on  a  central  stud.     Buckling  and  distor- 
tion of  the  discs  is  prevented  by  a  guard  which  limits  the  lift. 

57.  Pump-Valve  Seats  May  Be  Either  Forced  Or  Threaded 
Into  The  Valve  Decks  (Figs.  50  and  43).     Where  the  seat  is 


Stem- 


-Spring 


Taper  J!t- 

Peaned  Edge-'' 

FIG.  50. — Rubber  Valve  For  Low  Pressure 
Warm-  Or  Cold- Water  Service. 


FIG.  51. — Flat-Faced  Wing  Poppet  Valve. 


forced  into  the  deck,  the  hole  in  the  deck  is  bored  to  a  very 
slight  taper,  and  the  cylindrical  portion  of  the  seat  is  turned  to 
correspond.  When  the  seat  has  been  forced  in,  the  projecting 
edge,  E  (Fig.  50)  is  peaned  over  to  prevent  the  seat  from  work- 


SEC.  58] 


DIRECT-ACTING  STEAM  PUMPS 


47 


Discharge 
Outlet'. 


ing  out.     Where  the  seat  is  threaded  into  the  valve  deck,  the 
threads  are  turned  on  a  slight  taper  to  insure  a  tight  fit. 

58.  Flat-Faced   Bronze  Poppet  Valves  (Fig.  51)  are  used 
on  pumps  of  the  pot-valve  type  (Fig.  46)  for  high  pressures. 
The  vertical  movement  of  these  valves  is  guided  by  wings 
which  work  in  the  valve-seat  openings. 

59.  Three    Different  Methods  Of  Arranging  The  Valves 
Of   Horizontal   Double-Acting    Suction   Pumps    are   in  use: 
(1)   The  sets  of  discharge-  and  of  suction-valves  may  be  super- 
imposed one   above  the  other  (Fig.  52)  above  the  pump-barrel 
or  cylinder.     (2)   The  sets  of  suction-  and  discharge-valves  may 
be  arranged  (Fig.  53)  side 

by  side  above  the  pump- 
barrel  or  cylinder.  (3) 
The  discharge-valves  may 
be  located  (Fig.  1)  above 
the  pump-barrel  or  cylin- 
der and  the  suction-valves, 
below. 

Arrangement  (1)  is 
commonly  used  in  small 
low-  or  medium-pressure 
pumps.  It  admits  of 
easy  access  to  the  valves 
for  renewals  and  repairs. 
Its  disadvantage  is  that 
it  (Fig.  52)  requires  a 
reversal  of  the  flow  of 
water  through  the  pump. 
This  tends  to  a  dimin- 
ished pumping  capacity.  A  pump  having  this  arrangement 
is  termed  a  submerged-piston  pump.  Practically  all  small 
boiler-feed  pumps  (Sec.  198)  and  wet  vacuum  pumps  (Sec. 
353)  are  so  constructed. 

Arrangement  (2)  is  commonly  used  in  pumps  (Fig.  53) 
designed  for  high  pressures.  It  permits  a  structural  design 
which  is  conducive  to  great  strength. 

Arrangement  (3)  is  much  used  in  large  low-  or  medium- 


Pump 
Cylinder--' 


(        I    ^  Inlet- 

\\\\\\\\\\\\\\\W\\\N 


FIG.  52. — Medium-Pressure  Piston-Pump  With 
Suction  And  Discharge  Valves  Arranged  Above 
Pump-Barrel. 


48 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


pressure  pumps.     It  permits  the  water  to  pass  through  the 
pump  without  any  reversal  of  flow. 


Sucthn  Valve- ^ 


x.  -  -  -Discharge  Or  if  fee 
Valve  Pots 


Suction 
-Orifice 


Pump 
.--Barrel 


FIG.  53. — Outside-Packed-Plunger  High-Pressure  Pump  With  Suction-  And  Discharge- 
Valves  Arranged  Side  By  Side  Above  Pump-Barrel. 

60.  The  Total  Effective  Area  Of  Opening  Of  Each  Set  Of 
Suction-  And  Discharge -Values  In  A  Direct-Acting  Steam- 


FIG.  54. — Open  Position  Of  Flat-Disk 
Pump- Valve. 


FIG.    55.— Water-End   Of    Single   Direct- 
Acting  Steam-Pump  With  Air-Chamber. 


Pump  should,  for  low  speeds,  be  about  30  per  cent.,  and  for 
high  speeds,  about  50  per  cent.,  of  the  piston-  or  plunger  area. 


SEC.  61] 


DIRECT-ACTING  STEAM  PUMPS 


49 


NOTE. — THE  AREA  OF  OPENING  GIVEN  BY  A  FLAT  Disc  VALVE  (Fig. 
54)  is  the  annular  area  obtained  by:  Multiplying  the  lift,  L,  in  inches,  by 
the  diameter,  d,  in  inches,  and  by  3.14.  The  most  adaptable  valve- 
diameter  has  been  found  to  be  from  3  to  4  in.  The  lift  commonly  used 
is  about  3'^  in.,  regardless  of  the  water-diameter. 

EXAMPLE. — The  water-piston  of  a  high-speed  pump  is  of  10-in.  diame- 
ter. The  piston-rod  is  of  3-in.  diameter.  How  many  flat  disk  valves, 


A\\V\\\V\\\\\\\\\\\\\\WS 

Compressor  Discharge-Pipe--  • 
FIG.  56. — Apparatus  For  Replenishing  Air-Chamber  In  Discharge-Pipe  Of  Hydraulic 
Elevator  Pump  Under  800  Lb.  Pressure  Per  Sq.  In. 

each  of  3-in.  diameter  and  %-in.  lift,  are  required  for  each  set  of  suction- 
and  delivery-valves  in  this  pump? 

02    v/    Q  7854- 

SOLUTION. — The  effective  piston-area  =  (102  X  0.7854) — x~~~ 

=  75  sq.  in.  The  area  of  opening  of  each  valve  will  (Sec.  60)  be  0.25  X 
3  X  3.14  =  2.36  sq.  in.  Hence,  (Sec.  60)  75  X  0.5  -r-  2.36  =  15.9,  or, 
practically,  16  valves  are  required  in  each  set. 

61.  Air-Chamber s  are  often  connected  to  the    discharge- 
valve  chambers  (Fig.  55),  or  to  the  discharge  pipes  (Fig.  56),  of 


50 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


Both  Opening  To- 
ward Air- Chamber- 


Air-Chamber- 


-Gate  Valve 


direct-acting  steam-pumps.     The  function  of  an  air-chamber  is 
to  provide  a  cushion  for  the  discharged  water. 

EXPLANATION. — The  air  in  the  chamber,  C,  (Fig.  55),  is  compressed, 
during  discharge,  to  a  pressure  approximately  equal  to  the  pressure 
against  which  the  pump  is  working.  Thus,  it  forms  a  highly  elastic 
buffer  or  cushion.  When  the  piston  reaches  the  end  of  its  stroke,  the 
discharge  suddenly  ceases.  An  instant  elapses  before  the  opposite 
stroke  begins.  During  this  instant,  expansion  of  the  compressed-air 
in  C  tends  to  keep  the  discharged  water  in  motion.  Hence,  the  reacting- 
tendency  of  the  column  of  water  in  the  discharge  pipe  is  neutralized. 
Consequently,  where  the  air-chamber  is  of  the  proper  proportions,  no 
shock,  neither  to  the  piping  nor  to  the  pumping  mechanism  should  result 
therefrom. 

NOTE. — THE  AIR-CHAMBER  Is  A 
LESS  NEEDFUL  ACCESSORY  IN  DU- 
PLEX-PUMP SERVICE  THAN  IN  SIM- 
PLEX-PUMP SERVICE.  Duplex-pumps 
(Sec.  71)  have  continuous  piston- 
travels.  Hence,  with  such  pumps,  the 
discharge  of  water  is  approximately 
continuous.  For  high -pressure 
duplex-pumps  (Sec.  75)  and  those 
working  against  very  high  pressures, 
as  in  high-pressure  hydraulic  elevator 
service,  air-chambers  are,  neverthe- 
less, distinctly  necessary. 

62.  The  Height  Of  The  Water- 
Level  In  The  Air-Chamber  Of  A 
Pump  (Fig.  55)  should  not  ex- 
ceed one-fourth  the  height  of  the 
chamber.  With  small  slow- 
running  pumps,  working  against 
pressures  below  50  Ib.  per  sq. 
in.,  it  is  usual  to  rely  entirely 
upon  the  air-bubbles,  which  are 
entrained  with  the  suction- 


Discharge'' 

Stroke 

Suction  Stroke'' 
Plane  of  Suction-Valve  Deck' 


FIG.  57. — Snifter  For  Replenishing 
Air-Chamber  Of  Direct-Acting  Steam- 
Pump. 


water,  for  maintaining  the  requisite  volume  of  air  in  the 
chamber.  Where  pumps  run  at  high-speeds  and  against 
pressures  higher  than  about  50  Ib.  per  sq.  in.,  good  service 
requires  that  the  air-chambers  be  recharged  occasionally  by 
mechanical  means,  or  by  use  of  a  snifter.  The  snifter  (Fig. 
57)  may  be  operated  by  the  pump  itself.  It  is  suitable  for 


SEC.  63]  DIRECT-ACTING  STEAM  PUMPS  51 

pumps  running  against  pressures  up  to  about  200  Ib.  per  sq.  in. 
It  can  be  used  only  where  the  pump  has  a  suction-lift. 

EXPLANATION. — The  snifter  is  connected  to  the  pump-cylinder  at  a 
point,  P,  (Fig.  57)  between  the  suction-  and  discharge  valve-decks. 
When  the  valves  V  and  Vi  are  opened,  water  is  forced,  during  the 
head-end  discharge-stroke  of  the  pump-piston,  into  the  snifter-cylinder, 
S.  The  air  in  S  is  thus  dislodged  and  forced  into  the  air-chamber,  A, 
through  the  check-valve  C.  During  the  corresponding  suction-stroke, 
the  water  in  S  is  drawn  back  into  the  pump  cylinder.  Thus  the  snifter- 
cylinder  is  again  filled  with  air  through  the  check-valve  C\. 

The  flow  through  valve  V\  should  be  throttled  on  the  suction-stroke 
to  prevent  all  of  the  water  from  being  drawn  from  cylinder  S.  The 
purpose  of  this  is  to  retain  a  column  of  water  in  S  to  act  as  a  piston  for 
driving  the  air  through  check-valve  C.  Valve  V\  should  be  so  ma- 
nipulated as  to  establish  a  regular  pulsation,  within  the  length  of  the 
glass  gage,  of  the  water-level  in  S. 

63.  Air-Chamber  Charging -Apparatus  For  Pumps  Working 
Against  Very  High  Pressures,  usually  depend  (Fig.  56)  for 
their  effectiveness,  upon  the  tendency  of  particles  of  com- 
pressed- air  to  percolate  through  masses  of  water. 

EXPLANATION. — Gate  valve  V  (Fig.  56)  being  closed  and  Vz  opened, 
the  air  compressor,  C,  is  started.  Gate  valve  V\  is  then  opened  to 
permit  the  water  in  the  reservoir,  R,  to  be  blown  out,  after  which  it  is 
closed.  When  the  pressure  within  the  reservoir  reaches  the  limit  of  the 
compressor's  capacity  for  compression,  which  may  be  about  75  Ib.  per 
sq.  in.,  valve  V2  is  closed  and  V  is  opened.  Water  then  passes  through 
the  connecting-pipe  and  gradually  fills  reservoir  R.  Coincidentally,  the 
compressed  air,  thus  displaced,  bubbles  through  the  water  in  the  con- 
necting-pipe and  upward  through  the  mass  of  water  in  the  lower  part 
of  the  air-chamber,  A.  The  gage-cocks,  G,  are  used  to  determine  the 
approximate  height  of  the  water  in  the  air-chamber,  A. 

64.  The  Ratio  Of  Air-Chamber  Volume  To  Volume  Of 
Water -Piston  Displacement  In  Direct-Acting  Steam  Pumps 

may,  for  ordinary  rates  of  speed,  be  about  as  follows:  (1)  For 
simple  pumps  (Fig.  57)  from  2  to  3.5.  (2)  For  duplex  pumps 
(Fig.  58)  from  1  to  2.5.  The  air-chamber  volume  of  a  pump 
for  high-speed  service  (Fig.  25)  may  be  from  5  to  6  times  the 
volume  of  piston  displacement. 

65.  Vacuum  Chambers  (Figs.  59  and  60)   are  sometimes 
attached  to  the  suction-pipes  of  direct-acting  steam-pumps. 
The  function  of  a  vacuum-chamber  is  to  insure  that  the  pump- 


52 


STEAM  POWER  PLANT  AUXILIARIES 


[Drv.  2 


Air 
Chamber 


FIG.  58. — Showing  Height  Of  Water 
In  Vacuum-Chamber  At  Instant  Of 
Piston-Reversal. 


Air 
Chamber-  -  - , 


Discharge 
,- Outlet 


Suction 
Inlet  to 
Pump 


FIG.  59. — Vacuum-Chamber  Con- 
nected To  End  Of  Suction  Pipe  Of 
Direct-Acting  Steam-Pump. 


Air 
Chamber 


Suction -Pipe- 
FIQ.  60. — Special  Form  Of  Vacuum-Chamber. 


SEC.  66] 


DIRECT-ACTING  STEAM  PUMPS 


53 


cylinder  be  completely  filled  with  water  at  each  reversal  of 
the  piston-stroke.  It  also  provides  an  air-cushion  for  the 
column  of  water  in  the  suction-pipe  when  the  movement  of 
the  water  is  suddenly  arrested,  due  to  the  momentary  stoppage 
of  the  piston  at  the  end  of  each  stroke. 

EXPLANATION. — During  the  piston-stroke  the  air  (Fig.  58)  in  the 
vacuum-chamber  tends  (Fig.  61)  to  expand.  Therefore,  if  the  current  of 
water  in  the  suction-pipe  is  insufficient  to  completely  fill  the  space  behind 
the  piston,  a  portion  of  the  water  standing  above  the  plane,  XY,  of  the 
suction-inlet  is  forced  into  the  cylinder.  Thus,  the  cylinder  will  be  full 
of  water  when  the  piston-stroke  is  reversed.  When  the  flow  of  water 
(through  the  suction  valves)  momentarily  ceases  at  the  end  of  the  stroke, 


Suction 
Pipe--- 


FIG.  61. — Showing  Height  Of  Water  In  Vacuum-Chamber  During  Progress  Of  Piston 

Stroke. 

the  momentum  of  the  moving  column  in  the  suction-pipe  is  expended  in 
compressing  (Fig.  58)  the  air  in  the  vacuum  chamber.  Thus  the  shock 
that  might  otherwise  attend  abrupt  stoppage  of  the  flow  is  avoided. 

66.  Direct-Acting  Steam-Pumps  May  Be  Classified,  With 
Reference  To  Their  Cylinders,  as  follows:  (1)  Single  or 
simplex  pumps.  (2)  Duplex  pumps.  A  simplex  pump  (Fig. 
62)  has  one  steam-cylinder  and  one  water-cylinder.  A 
duplex  pump  (Figs.  63  and  64)  has  two  steam  cylinders  and 
two  water  cylinders.  It  comprises,  in  effect,  two  single 
pumps,  A  and  J5,  (Fig.  63)  placed  side  by  side,  drawing  water 
through  a  common  suction-pipe,  /S,  and  discharging  into  a 
common  delivering  chamber,  C,  and  pipe  D. 


54 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


FIG.  62. — Longitudinal  Sectional  Elevation  Of  Burnham  Direct-Acting  Simplex  Steam- 
Pump. 


-left-Hanoi  Siate     ..-Abutments 


TL..— P/5/OA7- 


Pis  ton -Roof s 


\    Delivery 
\  I  "Plan  View  Of  Valve  Gear,  Slide-Valves  Ceniral     ';.-Choimber 


Elevation  Of  Valve-Gear,  Rocker-  Ju-rns  Perpendicular 
Fia,  03,— Plan  And  Elevation  Of  Valve  Gear  Of  Duplex  Steam  Pump. 


SEC.  67] 


DIRECT-ACTING  STEAM  PUMPS 


55 


NOTE. — EACH  STEAM- VALVE  OF  A  DUPLEX  PUMP  Is  ACTUATED  BY 
THE  OPPOSITE  PISTON-ROD.  The  reciprocative  motion  of  the  piston- 
rods,  #x  and  #2,  (Fig.  63)  is  transmitted  to  the  slide-valves,  Vi  and  F2, 
through  a  system  (Fig.  65)  of  oscillating  rocker-shafts  and  arms. 


67.  The  Steam-Valve  Gears  Of 
Simplex-Pumps  are  (Figs.  62  and 
66)  variously  constructed.  With 
all  forms  of  such  gears,  however, 
the  main  valve  for  admitting  steam 
to  the  cylinder  and  releasing  it 
therefrom,  is  operated  by  direct 
steam-pressure.  The  valve  is  thus 
said  to  be  steam-thrown. 


Steam  Chest, 

nm 


,'Rocker-Sfancf 


; 

, 

-s 

?- 

/ 

_]* 

. 

V 

• 

\ 

^•.. 

n 

Short 


FIG.  64. — Sectional  Elevation  Of 
One  Side  Of  Vertical  Duplex  Steam 
Pump  For  Boiler  Feed  Service. 


../Cross/?e<YC/  Spools 


FIG.  65.  —  End-View  Of  Steam-  Valve-Actuat- 
ing-Mechanism  Of  Duplex  Pump. 


EXPLANATION. — At  the  beginning  of  the  inboard  stroke  (Fig.  62), 
main  steam-port  E  is  covered  by  the  piston,  P.  Enough  steam  to  give 
the  piston  an  easy  start  passes  in  behind  it  through  pre-admission  port 
GI  (Figs.  62  and  67).  When  the  piston  moves  far  enough  to  uncover 
port  E,  it  receives,  through  the  valve-port  F2  (Figs.  62  and  67)  and  main 
steam-port  E,  the  full  steam  pressure.  It  then  moves  at  normal  speed 
until  it  covers  main  steam  port  F  (Fig.  62).  By  this  covering  of  port  F, 
the  exhaust  steam  ahead  of  the  piston  is  trapped  in  the  inboard  end  of 


56 


STEAM  POWER  PLANT  AUXILIARIES 


[Dw.  2 


tubrKottor  Connection-- 


Steam 

Chest-... 

Plunger 


----Steam 
Chest 


the  cylinder.     The  exhaust  steam  thus  forms  a  cushion,  against  which  the 
piston  makes  an  easy  stop. 

During  the  inboard  stroke,  the  actuating  lever,  A,  is  shifted  to  the 
opposite    angular  position,  as  indicated   (Fig.   62)   by  the  dotted  lines. 

The  toe  of  the  actuating  lever 
thus  strikes  tappet-block  K  and 
shifts  the  auxiliary  slide-valve, 
H  (Fig.  67)  far  enough  to  the 
left  to  open  communication 
between  auxiliary  steam-port 
C2  and  auxiliary  exhaust-port 
R.  Coincidentally,  the  auxiliary 
valve,  H,  will  admit  live-steam, 
through  auxiliary  port  Ci,  to 
the  right-hand  end  of  chest- 
piston,  M  (Figs.  62  and  67). 
This  will  cause  the  chest-piston, 
which  engages  with  and  shifts 
the  main  slide-valve  D,  to  move 
instantly  to  the  left.  Thus  the 
main  steam-port  F  will  be 
brought  to  coincide  with  the 
drilled  ports,  V\t  in  the  main 
slide-valve.  Preadmission  port 
Gz  will  likewise  be  open.  Steam  will  thus  be  admitted  to  the  right- 
hand  end  of  the  cylinder  for  a  reversal  of  the  piston-stroke. 

If  steam  is  admitted  to  the  steam  chest,  M,  (Fig.  66),  it  will  enter  the 
hollow  ends,  H,  of  the  steam-chest  plunger,  F,  and  issue  through  a  hole 

,-— :: Steam  Auxiliary  tori-s---     -N      Actuaf!ny.Leyer  p^-stud. 

•I .'Auxiliary  Exhaust  Port ,-— Auxiliary  Slide,  Valv&\ 


Reversing  Valves 

?W$8$^^ 

FIG.  66. — Sectional  Elevation  Of  Steam-End 
Of  Cameron  Direct-Acting  Simplex  Steam- 
Pump,  Showing  Inside-Operated  Valve 
Mechanism. 


FIG.   67. — Plan  Of  Steam- Valve  Gear  Of  Burnham  Direct-Acting  Simplex  Steam-Pump. 

in  each  end.  The  spaces  between  the  ends  of  the  plunger,  F,  and  the 
heads  of  the  steam  chest  will  thus  be  filled  with  steam.  Steam  will  also 
enter  the  cylinder  through  the  port  PI  and  drive  the  main  piston,  C, 
to  the  left.  When  the  main  piston  strikes  the  stem  of  the  reversing 


SEC.  68]  DIRECT-ACTING  STEAM  PUMPS  57 

valve  Rz  and  forces  this  valve  to  the  left,  the  steam  at  the  left-hand  end 
of  the  plunger,  F,  will  escape  through  the  port,  E2,  into  the  annular 
cavity  A2  and  thence  through  a  cored  passage  (not  shown)  in  the  cylinder 
casting  into  the  exhaust  cavity,  K.  The  balance  of  pressure  between  the 
two  ends  of  the  steam-chest  plunger,  F,  will  thus  be  destroyed.  Due  to 
the  preponderence  of  pressure  at  the  right-hand  end,  the  plunger  will  be 
instantly  thrust  to  the  left-hand  end  of  the  steam  chest.  The  slide  valve, 
D,  is  attached  to  the  plunger,  F.  Hence  it  will  likewise  be  shifted  to  the 
left.  Live  steam  will  then  enter  the  left-hand  end  of  the  cylinder  through 
the  port  P2,  while  the  spent  steam  in  the  right-hand  end  will  be  exhausted 
through  the  port  PI. 

Instantly,  when  the  main  piston,  C,  starts  on  the  return  stroke,  the 
reversing  valve,  R2,  will  be  closed  by  the  pressure  of  the  steam  which  is 
constantly  in  contact  with  it  through  the  dotted 'port  $2-  When  the 
main  piston  has  traveled  far  enough  to  the  right,  it  will  shift  the  reversing 
valve  Ri.  The  series  of  events  described  above  will  then  be  repeated 
at  the  right-hand  end. 

68.  The  Length  Of  Stroke  Of  A  Simplex  Pump  Having 
External  Valve  Gear  (Fig.  62)  depends  upon  the  adjustment 
of  the  auxiliary  slide-valve,  H,  (Fig.  67). 

EXPLANATION. — Prick-punched  shop-marks  on  the  tie-rod,  X,  which 
forms  the  bearing  for  the  piston-rod  guide  (Fig.  62)  indicate  the  extreme 
travel  of  the  piston  in  each  direction.  If  the  inboard  stroke  is  too  short 
it  may  be  lengthened  by  a  slight  shifting  of  the  tappet-block  L  (Fig. 
62),  along  the  valve-stem,  toward  the  right.  The  outboard  stroke  may 
likewise  be  lengthened  by  shifting  the  tappet-block  K  toward  the  left. 
These  adjustments  will  permit  the  actuating-lever,  A,  (Fig.  62)  to  oscil- 
late further  in  each  direction  before  striking  the  tappet-blocks.  Shifting 
of  the  auxiliary  valve  H,  (Fig.  67)  will  thus  be  delayed. 

If  the  piston-rod  guide  travels  very  close  to  the  marks,  the  piston  may 
hesitate  before  reversal  at  the  end  of  each  stroke.  Or,  it  may  sometimes 
hang  at  the  end  of  a  stroke.  When  this  occurs  the  tappet-blocks,  K 
and  L,  (Fig.  62)  should  be  shifted  closer  to  the  toe  of  the  actuating  lever. 

69.  Adjustment  Of  The  Steam-Valve  Of  A  Direct-Acting 
Duplex-Pump  (Fig.  63)  consists,  first,  in  plumbing  the  rocker 
arms  and  setting  both  valves  line-and-line  with  the  outer 
edges   of   the   steam-ports.     The   lost-motion,    or   clearance, 
between  the  valve-stem  collars  and  the  abutments  on  the 
backs  of  the  valve  is  then  divided  equally.     See  Sec.  73. 

70.  To  Determine  The  Requisite  Length  For  Either  The 
Steam-Valve  Rod  Or  Stem  Of  A  Duplex  Pump  proceed  as 


58 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


follows:  Place  the  valve-arm  plumb  (Fig.  68)  and  put  the 
valve  in  its  central  position.  The  valve  will  be  central  when 
its  outside  edge  at  each  end  coincides  with  the  outside  edge 
of  the  corresponding  steam  port.  If  the  valve  steam  is 
missing,  the  valve-rod  should  be  blocked  up  to  a  horizontal 
position.  The  length  of  the  missing  stem  will  then  be  given 
by  the  distance  A  (Fig.  68).  In  laying  off  this  distance,  the 
clearance,  C,  between  the  end  of  the  stem  and  the  wall  of  the 
steam-chest,  should  be  greater  than  the  steam-port  width, 
assuming  this  to  be  equal  to  the  maximum  displacement  of  the 
valve  from  its  central  position.  If  the  valve  rod  is  missing, 
the  stem  should  be  inserted  and  the  lost  motion,  L  and  LI 
accurately  adjusted.  The  length  of  the  missing  rod  will  then 
be  given  by  the  distance  B  (Fig.  68). 


Fia.  68. — Method  Of  Finding  Lengths  Of  Steam- Valve  Stems  And  Rods  Of 
Duplex  Pumps. 


71.  The  Function  Of  The  Valve-Stem  Lost-Motion  In 
Duplex  Direct  Acting  Steam-Pumps  (Fig.  63)  is  threefold: 
(1)  It  permits  adjustment  of  the  piston-stroke.  (2)  It  causes  a 
continuous  piston-travel.  (3)  It  prevents  the  pump  from  stopping 
in  a  position  from  which  it  cannot  be  started  by  admitting  steam 
to  the  steam  chest.  Continuous  piston-travel  is  secured  by 
preventing  simultaneous  reversals  of  the  piston-strokes. 
Assurance  against  a  dead-center  or  non-starting  position  is  due 
to  the  fact  that  when  either  steam-valve  covers  all  four  ports 
(Fig.  63)  the  opposite  valve  leaves  an  admission — and  an  ex-' 
haust-port  wide  open.  This  feature  of  the  duplex  pump 
renders  it  well-adapted  for  periodic  operation  (Figs.  69  and  70) 
under  governor  control. 


SEC.  71] 


DIRECT-ACTING  STEAM  PUMPS 


59 


*+a~m  *  «„/        Governor  Huo/faulic  Air- 

,-oT-ewm  Supply.       Pressure  Pipe-  Chamber    " 

Line  r    •„ 


Dlscha 
To  Sprfri 
Pipe  n're- 

System-,^ 

Check 


•-Fisher  Steam 

Pump  Governor      /Re//ef  Voi,Ye 


'  ''Auxiliary  Duplex  Pump 

FIG.  69. — Underwriters    Fire-Pump    Equipment    For    Connection    To    Sprinkler-Pipe 

System. 


Governor  Pipe  for  Operation  of  Governor  Unoter 
\          Gravity  Pressure  of  Water 

~  T?-K?ffi '  1 1  i-LLJ-LLLiii  mrmi 

B3Edtf3W 


^$$$^^ 

Fia.  70. — Governor-Controlled  Duplex  Pump  For  Water-Service  In  Buildings. 


60 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


EXPLANATION. — A  duplex  pump  (Fig.  63)  is  presumed  to  have  been 
correctly  adjusted  for  running.  The  steam-pistons,  PI  and  P2,  (Fig. 
71)  are  at  mid-stroke.  Likewise,  the  slide-valves,  Vi  and  F2,  are  at  mid- 
travel.  Piston  P2  actuates  valve  V\  through  the  long  rocker-arm, 


Exhaust  Port-^^  Admission  Port 
Center  Lines  of 
Piston  Rocfs- 


Centerlines  of 
Votive  Stems-- 


FIG.  71. — Pistons  And  Valves  In  Mid- 
Position. 


Admission  Port-- 
Lost Motion 


FIG.  72. — One  Steam-Valve  Shifted 
From  Central  Position.  Piston  P2 
Starts  Movement  To  Left. 


CenterLines  of 
Piston  Roots-,, 


Cross 
.-Head. 


pivoted  at  F\.  Piston  PI  actuates  valve  Vz  through  the  short  rocker- 
arm,  pivoted  at  F%.  The  rocker-arms  stand  perpendicularly  to  the  center- 
lines  of  the  cylinders.  The  lost  motion  of  each  valve-steam  is  equally 
divided  between  the  ends  of  the  valve. 

Valve  Vz  (Fig.  72)  has  been  moved 
so  that  the  steam-port  S  is  open  for 
admission  of  steam  to  the  cylinder. 
It  is  presumed  that  this  was  done 
before  the  steam-chest  covers  (Fig. 
63)  were  put  on.  If  the  valves  had 
been  left  as  in  Fig.  71,  the  pump  could 
not  start  when  steam  would  be  ad- 
mitted through  the  throttle-valve  in 
the  supply-pipe. 

Steam  has  entered  through  port  S 
(Fig.  72)  and  has  driven  piston  P  •>  in 
the  direction  of  the  arrow.  This 
piston  has  moved  just  far  enough  to 
take  up  the  lost-motion  at  the  right- 
hand  end  of  valve  V\.  Further  move- 
ment will  open  steam-port  Si. 

Piston  P2  (Fig.  73)  has  completed  a  stroke.  Coincidentally,  piston  PI 
has  traveled  far  enough  in  the  direction  of  the  arrow  to  shift  valve  Vz 
to  the  mid-travel  position,  where  it  is  on  the  edge  of  opening  steam-port 
Sz  for  the  return-stroke  of  piston  P2. 


FIG.  73.— Piston  P2  At  End  Of  In- 
board Stroke  And  On  Point  Of  Re- 
versal-Piston Pi  Approaching  End  Of 
Inboard  Stroke. 


SEC.  72] 


DIRECT-ACTING  STEAM  PUMPS 


61 


If  the  stem  of  valve  F2  had  less  lost  motion,  the  valve  could  not  have 
been  shifted  (Fig.  72)  far  enough  to  give  a  full  opening  at  port  S.  Hence 
valve  F2  would  have  been  moved  to  the  mid-travel  position  (Fig.  73) 
before  piston  P2  could  have  reached  the  end  of  the  cylinder-bore.  Thus  a 
short-stroke  would  have  resulted.  On  the  other  hand,  if  the  lost  motion 
were  greater  than  the  amount  shown  (Fig.  72),  valve  V2  could  not  reach 
mid-position  (Fig.  73),  and  thereby  close  ptfrt  S  against  admission  of 
steam  behind  piston  P2,  coincident  with  the  arrival  of  that  piston  at  the 
end  of  the  cylinder. 

Piston  PI  (Fig.  74)  has  finished  its  stroke.  Coincidentally,  it  has 
moved  V2  to  the  limit  of  its  right-hand  travel.  Steam-port  S2  has  thus 
been  fully  opened  for  admitting  steam  behind  piston  P2.  Piston  P2 
has,  therefore,  moved  far  enough  on  its  return  stroke  to  shift  valve  V\ 
to  the  edge  of  opening  steam-port  S3  for  a  reversal  of  piston  PL 


Exhaust  Port—  -,  Admission  Port-, 
Center  Lines  of 


Admission  Port--^ Exhaust  Port— .. 

Center  Lines  of 
Piston  Roo/s-^ 


Exhaust  Port-  •  •    Admission 


FIG.  74.— Piston  Pi  At  End  Of  Inboard 
Stroke  And  On  Point  Of  Reversal.  Piston 
Pz  Approaching  End  Of  Head-End  Stroke. 


Center  Lines  of 
Valve  Stems 

Exhaust  Port—^ 


FIG.  75.— Piston  Pz  At  End  Of  Head- 
End  Stroke  And  On  Point  Of  Reversal. 
Piston  Pi  Approaching  End  Of  Head-End 
Stroke. 


Piston  P2  (Fig.  75)  has  completed  its  return  stroke.  Coincidentally, 
piston  PI  has  traveled  far  enough  on  its  return  stroke  to  shift  valve  F2 
to  the  edge  of  opening  port  S  for  another  reversal  of  piston  P2. 

NOTE. — INCORRECT  ADJUSTMENT  OF  THE  VALVE-STEM  LOST-MOTION 
IN  DUPLEX  PUMPS  MAY  BE  A  SOURCE  OF  Loss.  When  the  pistons  do 
not  reach  the  limits  of  possible  travel,  they  must  make  many  more 
strokes  than  would  otherwise  be  required  to  do  the  same  amount  of  work. 
This  means  extra  consumption  of  steam  and  cylinder  oil,  and  extra  wear, 
particularly  of  the  water  valves. 

72.  The  Points  At  Which  The  Cross-Heads  Should  Be 
Secured  To  Duplex-Pump  Piston-Rods  may  be  determined  as 
follows:  The  packing  should  be  removed  from  the  piston-rod 


62 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  2 


stuffing-boxes  (Fig.  76)  and  the  glands  should  be  screwed  up 
tightly.  The  cylinder-heads  being  removed,  the  steam- 
pistons  should  be  pushed  up  solidly  against  the  center-heads. 
A  line,  A,  (Fig.  76)  should  then  be  scribed  on  each  rod  flush 
with  the  faces  of  the  water-end  glands  or  gland-nuts.  The 
heads  of  the  steam-cylinders  should  then  be  put  on.  The 


Steam- 


,- -Center- Heaof 
'.Striking  Point 


Nut 


FIG.  76. — Marking  Center-Head  Striking 
Point. 


FIG.  77.— Marking  Cylinder-Head  Strik- 
ing-Point. 


••Steam  End       Rocker  Arm-. 


Water  End- 


heads  of  the  water-cylinders  being  removed,  the  steam-pistons 
should  be  pushed  up  solidly  (Fig.  77)  against  the  cylinder- 
heads.  A  line,  B,  (Fig.  77)  should  then  be  scribed  on  each 
rod  flush  with  the  faces  of  the  steam-end  glands  or  gland-nuts. 
The  scribed  lines,  A  and  B,  (Fig.  78)  thus  establish  the  striking- 
points.  The  lines  should  be  prick-punched  to  make  them 
discernible  for  future  reference.  By  shifting  the  pistons  until 

A  and  B  become  equally  dis- 
tant (Fig.  78)  from  the  glands, 
the  pistons  will  be  placed  ex- 
actly at  mid-stroke.  The 
crossheads  should  then  be 
slipped  along  the  rods  until 
the  rocker-arms  (Fig.  78) 
stand  plumb.  The  cross- 
heads  may  then  be  clamped 
to  the  rods. 


FIG.  78.  —  Striking-Points  Equally 
Spaced,  Rocker-Arm  Plumb,  Crosshead  In 
Correct  Position. 


NOTE. — The  crossheads  of  new  duplex  pumps  are,  generally,  so  secured 
to  the  rods  as  to  preclude  possibility  of  error  in  restoring  the  crossheads 
should  they  at  any  time  be  temporarily  removed.  The  operation  of 
finding  the  striking  points  (Sec.  72)  is,  however,  necessary  where  the 
piston  rods  of  an  old  pump  have  been  renewed. 

73.  The  Correct  Amount  Of  Valve-Stem  Lost-Motion  In 
Duplex -Pumps  depends  upon  the  service  in  which  the  pump 


SEC.  74] 


DIRECT-ACTING  STEAM  PUMPS 


63 


is  to  be  used.  Pumps  designed  to  run  at  high  speeds  (Sec.  28) 
require  considerably  less  valve-stem  lost-motion  than  do 
pumps  for  slow-speed  service.  Generally,  lost-motion  (Fig. 
68),  at  each  end  of  the  valve,  equal  to  about  one-third  of  the 
admission-port  width  will  suffice  for  ordinary  service. 


NOTE. — THE  VALVE-STEM  LOST-MOTION 
DUPLEX-PUMPS,  as  those  in  boiler- 
feed  and  elevation  service,  should 
be  such  that  each  piston  will  travel 
nearly  full  stroke  before  shifting 
the  opposite  slide-valve  to  the 
admission  edge  of  the  steam-port. 


Vctlve  5fem~, 


IN      SLOW-RUNNINQ- 


Aotjustlny  Nuts 

•  •Sliding  Block 


---Votive  Arm 


Slide  Voi/ve-:  Valve  Seat- 

Fid.  79. — Rigid  Valve-Stem  Con- 
nection Of  Duplex-Pump  Slide- Valve- 
Lost-Motion  Provided  Externally. 


Croc/te- 


FIG.  80. — Mechanism  For  Outside  Ad- 
justment Of  Lost  Motion  In  Duplex-Pump 
Valve  Gear. 


NOTE. — THE  VALVE-STEMS  ARE  OFTEN  RIGIDLY  ATTACHED  (Fig.  79) 
to  the  slide-valves.  In  such  cases  a  link  mechanism  (Fig.  80),  with  sliding 
blocks  and  tappets,  is  provided  for  adjusting  the  lost-motion  outside  the 
steam-chest. 


.-Cushion  Votive. 


'-Compression  Space 
I-  Cross-Section   on  Line  X-Y  E'Lonoj  i  t  u  ol  i  rial     Section 


By  Pass- 
Section 
FIG.  81. — How  Duplex-Pump  Pistons  Are  Steam-Cushioned. 


74.  Compression-Space  In  The  Steam-Cylinders  Of  Duplex 
Pumps  is  the  volume  of  cylinder-space  S,  (Fig.  81)  in  front 
of  the  piston,  plus  the  volume  of  space  in  the  admission-port, 
A,  at  the  instant  the  piston  has  completely  closed  the  corres- 
ponding exhaust-port  E. 


64 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


EXPLANATION. — The  piston  P,  (Fio.  81)  has  reached  a  position  in  its 
travel  wherein  it  prevents  escape  of  steam  through  the  exhaust-port 
E.  Coincidentally  the  slide-valve,  D,  covers  the  admission-port  A. 
Hence  the  piston  will  be  cushioned  in  its  further  progress  by  compressing 
the  steam  ahead  of  it  in  the  space  S. 

NOTE. — Large  duplex-pumps  are  equipped  with  cushion-valves,  G, 
(Fig.  81)  for  adjusting  the  cushioning  effect  of  steam  in  the  compression 
spaces.  This  is  done  by  controlling  the  flow  of  steam,  through  the  by- 
pass, B,  from  the  admission-port  A  to  the  exhaust-port  E. 

75.  The  Relative  Merits  And  Demerits  Of  Simplex  And 
Duplex  Pumps  may  be  summarized  as  follows:  (1)  The  flow 
of  water  in  both  the  intake-and  discharge-pipes  of  a  simplex  pump 


Junction  Pipe... 


;-Chest  Piston-  -  - 

•  •  Live  Steam 
'   Connection-, 


-Slide  Valve 

Stem  of  Auxiliary  Valve  which 
Admits  Steam  to  Throw  Chtst 
P/ston-P* 


^ 

Connection  through  which  Chest  Pfston-P,  •               Vibrating  Rod  which  Actucftes; 
is  Operated  by  Chest  Piston-P •'  AuOHty  Valve-Stem - 

FIG.  82. — Sectional  Elevation  Of  Steam  Cylinders  Of  A  Burnham  Compound  Simplex 

Pump. 

must  cease  during  piston-reversal.  The  water  hammer  which 
tends  to  result  therefrom  may,  however,  be  prevented  or 
modified  by  using  (Sees.  61  and  65)  air  and  vacuum  chambers. 
(2)  With  a  duplex  pump  (Sec.  71)  the  flow  of  water  is  prac- 
tically continuous.  (3)  The  piston  of  a  simplex  pump  travels 
the  maximum  set  distance  during  each  stroke.  The  length  of  the 
stroke,  after  being  fixed  by  adjustment  of  the  auxiliary  steam 
valve  (Sec.  68),  continues  constant  regardless  of  the  retarding 
tendency  of  piston-,  rod-,  and  cylinder-friction.  (4)  The 
pistons  of  a  duplex  pump  may  short-stroke.  Short-stroking 
may  be  due  to  the  retarding  effect  of  friction  between  the 
pistons  and  cylinders  and  in  the  piston-rod  stuffing-boxes. 


SEC.  76] 


DIRECT-ACTING  STEAM  PUMPS 


65 


(5)  The  simplex  pump  uses  less  steam  for  the  same  amount  of 
work  than  does  the  duplex  pump.  This  is  due  to  smaller 
clearance  spaces  in  the  steam  cylinder. 

NOTES. — SIMPLEX-PUMPS  ARE  WELL-ADAPTED  As  VACUUM-  AND 
AIR-PUMPS  in  connection  with  surface-condensers.  This  is  due  to  their 
comparatively  small  clearance  spaces  and  immunity  from  short-stroking. 

DUPLEX-PUMPS  ARE  WELL-ADAPTED  FOR  HIGH-PRESSURE  SERVICE. 
They  are  also  preferable  where  either  a  very  high  or  a  very  slow  velocity 
of  flow  is  required.  This  is  due  to  their  practically  continuous  action. 

FIRE-INSURANCE  UNDERWRITERS  require  that  DIRECT-ACTING  STEAM 
FIRE  PUMPS  (Fig.  25)  be  of  the  duplex  type.  These  pumps  are  com- 
monly connected  to  sprinkler-pipe  fire-systems.  In  such  cases  auxiliary 
duplex  pumps,  A,  (Fig.  69)  are  provided  for  making  up  the  leakage  from 
the  sprinkler  system  and  maintaining  a  constant  pressure  therein. 


.'Live  Steam 
Inlet 


Exhaust  Outlet  to  Feetfwater 
Heater,  Condenser  or  Atrmsphere 


FIG.  83. — Sectional  Elevation  Of  Steam  Cylinders  Of  A   Compound    Duplex-Pump. 

NOTE. — LARGE  DIRECT-ACTING  STEAM  PUMPS  ARE  OFTEN  BUILT 
WITH  COMPOUND  STEAM-CYLINDERS  (Figs.  82  and  83).  This  is  done  to 
economize  their  steam  consumption.  See  the  Author's  STEAM  ENGINES. 

76.  The     Steam-Piston    Areas     In    Boiler-Feed    Pumps 

(Sec.  28)  are  usually  from  about  two  to  three  times  the  water- 
piston  areas.  In  boiler-feed  service  the  total  head  and  the 
available  steam-pressure  are  practically  equal.  A  large 
excess  of  steam-piston  area  is,  however,  provided  as  a  safety 
precaution.  It  conduces  to  prompt  starting  of  the  pump. 

77.  Selection  Of  A  Direct-Acting  Steam  Pump  For  Boiler- 
Feed  Service  is  based  upon  two  main  factors:  (1)  The  steam- 
ing capacity  of  the  boilers  to  be  fed.     (2)  A  proper  rate  of 


66  STEAM  POWER  PLANT  AUXILIARIES  [Div.  2 

piston  travel.  The  pump  must  be  large  enough  to  deliver, 
while  running  at  a  moderate  speed,  the  maximum  quantity 
of  water  that  can  be  evaporated  in  the  boilers.  It  is  conve.,- 
tionally  assumed  that  these  conditions  are  fulfilled  by  selecting 
a  pump  that  will  deliver  45  Ib.  of  water  per  hour  per  boiler 
horse  power  while  running  at  one-half  the  rated  normal  speed 
of  the  pump. 

78.  Pump  Managment  is  discussed  in  the  following  notes. 
Although  these  directions  are  included  in  this  Div.  on  DIRECT 
ACTING  STEAM  PUMPS  many  of  the  suggestions  apply  with 
equal  weight  to  pumps  of  any  type.  This  material  is  quoted 
from  the  COAL  MINER'S  POCKETBOOK: 

ALL  PUMPS,  WHEN  NEW,  SHOULD  BE  RUN  SLOWLY  until  the  parts  have 
become  thoroughly  adjusted  to  their  bearings,  when  the  speed  may  be 
increased.  Because  a  new  pump  works  stiffly  is  no  cause  for  alarm,  for, 
while  a  machinist  can  properly  construct  the  parts,  he  cannot  always 
forsee  the  strains  caused  by  the  action  of  the  pump,  when  the  parts  are 
assembled  and  which  require  certain  adjustments  after  the  pump  is  at 
work.  By  running  the  pump  slowly  with  the  parts  properly  lubricated 
and  making  such  adjustments  as  may  be  necessary,  stiffness  will  gradu- 
ally disappear  and  the  highest  efficiency  of  the  pump  will  then  be  at- 
tained, provided  other  matters  on  which  the  pump's  action  depend  have 
received  proper  attention. 

THE  CAUSES  THAT  AFFECT  A  PUMP,  IMPAIR  ITS  EFFICIENCY,  AND 
PREVENT  IT  FROM  PERFORMING  ITS  FULL  DUTY  are:  (1)  wear;  (2)  the 
improper  adjustment  of  valves,  valve  stems,  and  levers;  (3)  the  improper 
packing  of  plungers  and  stuffing  boxes;  (4)  drawing  up  the  stuffing-box 
glands  too  tightly;  (5)  lost  motion  due  to  permitting  the  working  parts  to 
wear  and  not  adjusting  them  to  the  new  conditions;  (6)  accumulations  of 
foreign  matter  under  the  valves  or  in  the  strainer;  (7)  broken  valves  and  valve 
springs',  (8)  leakage  in  valves;  (9)  taking  air  in  the  suction  pipe;  (10) 
clogged  or  broken  discharge  pipes;  and  (11)  the  use  of  poor  gaskets. 

MANY  PUMPS  ARE  CAPABLE  OF  A  LARGER  CAPACITY  THAN  Is  OB- 
TAINED BY  THE  Low  SPEED  AT  WHICH  THEY  ARE  OPERATED,  but  it  is 
important  that  such  pumps  be  run  continuously,  as  any  serious  interrup- 
tion in  pumping  might  cause  trouble  elsewhere.  It  is  customary,  there- 
fore, to  keep  on  hand  a  supply  of  duplicate  valves,  moving  parts,  and 
packing,  in  order  that  when  it  becomes  necessary  to  make  repairs  they 
may  be  made  without  great  loss  of  time. 

A  COMMON  CAUSE  OF  PUMPS  REFUSING  To  WORK  PROPERLY  Is  DUE 
To  THEIR  TAKING  AIR  BELOW  THE  SUCTION  VALVES.  Small  leaks  will 
cause  the  piston  to  jump  owing  to  the  water  not  entering  through  the 
suction  valves  soon  enough  to  fill  the  entire  chamber.  This  trouble 


SEC.  78]  DIRECT-ACTING  STEAM  PUMPS  67 

may  be  remedied  by  making  all  joints  in  the  suction  pipe  and  between 
the  pipe  and  the  pump  air-tight.  Leaks  may  sometimes  be  detected  by 
the  hearing  or  by  the  flame  from  a  candle  being  drawn  toward  the  hole. 
If  the  leaks  are  small  and  not  at  the  pipe  joints,  a  coat  of  asphalt  paint 
may  stop  them;  if  large,  they  should  be  drilled  larger,  the  hole  threaded, 
and  a  screw  plug  inserted.  If  the  leak  is  at  the  joint  between  two  pipes, 
the  pipes  should  be  uncoupled  and  screwed  together  again,  using  graphite 
pipe  grease  for  a  lubricant.  Or,  if  the  joint  is  a  flanged  one,  a  new  gasket 
should  be  placed  between  the  flanges,  and  the  pipes  lined  up  before  the 
bolts  are  tightened. 

SOMETIMES,  A  PUMP  FAILS  To  CATCH  THE  WATER  WHEN  STARTED 
OWING  To  LEAKAGE  OF  THE  VALVES  IN  THE  SUCTION  CHAMBER.  The 
trouble  may  be  caused  by  the  valve  and  the  valve  seat  being  corroded; 
by  chips  or  gravel  getting  under  the  valves  and  preventing  them  from 
seating  properly;  or  by  the  valves  and  seats  becoming  worn  so  that  leak- 
age cannot  be  prevented  without  changing  the  parts. 

MANY  PUMPS  WILL  NOT  RAISE  WATER  IN  THE  SUCTION  PIPE  WHEN 
EMPTY,  OWING  To  THE  PUMP  HAVING  BEEN  IDLE  FOR  SOME  TIME, 
but  will  continue  to  draw  water  after  once  being  started.  In  such  cases, 
it  is  necessary  to  prime  the  pump,  by  which  is  meant  filling  the  suction 
pipe  and  part  of  the  suction  chamber,  if  there  is  one,  and  in  some  cases, 
also,  the  pump  barrel,  with  water,  so  that  the  pump  may  start  under 
conditions  similar  to  those  under  which  it  must  work.  To  prime  the 
pump,  open  the  cock,  or  valve,  in  the  priming  pipe  and  allow  water 
from  the  column  pipe  to  flow  down  into  the  suction  pipe  and  the  pump. 
When  these  are  full,  the  valve  is  again  closed  and  the  pump  is  ready  to 
start. 

PUMPS  SOMETIMES  FAIL  To  RAISE  WATER  WHEN  THE  FULL  HEAD  Is 
RESTING  ON  THE  VALVES  IN  THE  DISCHARGE  CHAMBER.  This  may  be 
due  to  air  accumulating  between  the  suction  and  the  discharge  decks, 
which  air  is  compressed  and  expanded  by  the  motion  of  the  plunger. 
Air  valves  should  be  provided  in  the  water  cylinder  to  allow  this  confined 
air  to  escape.  Violent  jarring  and  trembling  often  occur  if  the  discharge 
chamber  is  not  provided  with  either  an  air  chamber,  where  the  lift  is 
not  above  150  ft.,  or  with  an  alleviator,  for  lifts  above  that  distance. 
This  jarring  is  due  to  the  column  of  water  in  the  discharge  pipe  coming  to 
rest  suddenly  between  strokes  and  having  to  be  again  put  in  motion. 

IN  CASE  THE  PUMP  COLUMN  Is  FILLED  WITH  WATER  AND  THE  PUMP 
Is  STOPPED,  THE  WATER  WILL  RUN  BACK  THROUGH  THE  PUMP  IF  THE 
FOOT- VALVE  Is  NOT  TIGHT.  To  prevent  this,  a  gate  valve  or  a  check- 
valve  is  placed  a  short  distance  from  the  pump  in  the  column  pipe.  A 
gate  valve  wears  less  than  does  a  check-valve,  and  presents  no  obstruction 
to  the  flow  of  water  when  the  valve  is  open.  This  valve  is  useful  in  the 
column  pipe  to  maintain  the  pressure  off  the  valves  when  the  pump  is  not 
at  work,  and  also  for  keeping  water  from  running  back  into  the  pump 
chamber  when  the  valves  are  being  repaired. 

WHEN  STARTING  COMPOUND  PUMPS,  the  steam  pressure  on  the  high- 


68  STEAM  POWER  PLANT  AUXILIARIES  [Div.  2 

pressure-cylinder  piston  is  not  always  sufficiently  powerful  to  mov*  the 
plungers  against  the  resistance  of  the  water  in  the  discharge  pipe.  But, 
by  opening  the  gate  valve  in  the  by-pass  piping,  the  pressure  on  the 
plungers  is  relieved  for  a  sufficient  number  of  strokes  to  allow  the  steam  to 
reach  the  low-pressure  piston,  when  the  combined  force  of  the  two  pistons 
will  do  the  work.  The  by-pass  pipe  can  then  be  closed. 

VALVES  IN  THE  STEAM  END  SOMETIMES  WEAR  UNEVENLY  OR  THEIR 
STEMS,  BY  CONTINUAL  ACTION  WEAR  AND  CAUSE,  LOST  MOTION,  thus 
causing  a  back  pressure  and  irregular  action.  Anything  wrong  in  the 
steam  end  can  usually  be  determined  by  the  irregular  exhaust,  but  even 
this  may  be  deceptive  in  case  the  water-end  valves  are  leaking.  If  the 
steam  valves  are  suspected,  the  steam  chest  cover  may  be  raised  for  their 
inspection,  but  the  valves  should  not  be  disturbed  until  it  has  been  deter- 
mined, by  moving  the  water  piston  backwards  and  forwards  several 
times,  that  they  do  not  open  and  close  properly.  The  trouble  may  be  in 
the  levers  or  toggles  that  throw  them.  If  so  the  correcting  adjustments 
may  be  properly  made  without  disturbing  the  valves.  In  many  duplex 
pumps,  there  are  very  slight  differences  between  the  two  sides,  and  the 
amount  of  the  lost  motion  (Sees.  71  and  73)  between  the  valve  stem 
and  the  valve  should  be  carefully  adjusted.  Too  little  lost  motion  will 
cause  short  stroking,  while  too  much  will  allow  the  pistons  to  strike  the 
heads.  The  adjustment  requires  skill. 

SOMETIMES,  THE  VALVE  SEAT  OR  THE  VALVE  HAS  SOFT  SPOTS  THAT 
WEAR  FASTER  THAN  THE  REMAINDER  OF  THE  VALVE  AND  SEAT. 
Through  these  slight  depressions,  steam  will  blow  and  cut  both  valve  and 
seat  if  attention  is  not  given  them;  back  pressure  will  then  seriously 
interfere  with  the  working  of  the  pump.  If  the  defect  is  in  the  valve,  a 
new  one  can  take  its  place.  But  the  valve  seat,  if  a  part  of  the  steam 
cylinder,  will  require  an  entirely  new  cylinder,  and  hence  it  is  economy  to 
scrape  the  seat  until  the  depressions  are  removed.  A  try  plate  made  of 
steel  having  a  perfectly  level  surface  is  covered  with  chalk  and  carefully 
rubbed  over  the  valve  seat.  The  elevations  will  have  chalk  on  them,  the 
depressions  will  not.  The  elevations  are  scraped  with  a  chisel  made  of 
the  best  steel  until  they  are  worn  down  so  that  chalk  sticks  to  every  part 
of  the  seat  alike.  The  valve  is  treated  in  the  same  way  if  it  can  be  done 
without  too  much  expense.  The  valve  and  the  valve  seat  when  remov- 
able should  be  sent  to  the  shop  to  be  reground. 

THE  FIRST  STEP  AFTER  A  PUMP  HAS  BEEN  ERECTED  Is  To  CLEAN  OUT 
THE  STEAM  PIPING.  In  order  that  this  may  be  done  without  carrying 
foreign  matter  into  the  pump,  the  piping  is  left  disconnected  from  the 
pump  and  steam  at  full  boiler  pressure  is  allowed  to  blow  freely  through 
the  piping  and  valves  for  a  few  minutes.  Steam  is  then  shut  off  and  the 
piping  is  connected  to  the  pump. 

THE  NEXT  STEP  Is  To  BLOW  OUT  THE  STEAM  CYLINDERS.  To  do  this, 
the  cylinder  heads  should  be  put  on,  leaving  the  pistons  and  valves  out  of 
the  cylinders.  The  stuffing  boxes  should  be  closed,  which  is  most 
conveniently  done  by  placing  a  piece  of  board  between  the  stuffing  box 


SEC.  78]  DIRECT-ACTING  STEAM  PUMPS  69 

and  the  reversed  gland  and  then  setting  up  the  nut  on  the  stuffing  box 
studs.  When  the  gland  is  drawn  home  by  a  nut  outside  of  it,  a  circular 
piece  of  pine  board  may  be  placed  between  the  end  of  the  gland  and  the 
inside  of  the  nut  in  order  to  close  the  opening  through  which  the  piston 
rod  passes.  Steam  may  now  be  turned  on  the  main  steam  pipe  leading  to 
the  pump;  by  opening  the  throttle  valve  wide  at  short  intervals. 
Thereby  the  sand  and  scale,  in  the  ports  and  other  passages  and  spaces  of 
the  steam  end,  can  be  blown  out.  After  the  cylinders  have  been  blown 
out,  the  heads  and  covers  should  be  removed  and  all  foreign  matter 
blown  into  the  corners  and  chambers  of  the  cylinders  removed  by  hand. 
The  pistons,  valves,  cylinder  heads,  and  other  covers  can  then  be  put  in 
place.  The  blowing  out  of  the  pipes  and  cylinders  after  erection  is  often 
neglected  or  but  imperfectly  done,  with  serious  consequences  to  the  machine. 
It  cannot  be  too  thoroughly  done,  particularly  in  pumps  of  the  type  in 
which  the  steam  ports  and  exhaust  ports  are  on  top,  for  in  this  construc- 
tion the  sand  and  grit  are  deposited  in  the  bottom  of  the  cylinder  for 
the  piston  to  ride  on. 

THE  PACKING  OF  ALL  RODS  AND  STEMS  Is  THE  NEXT  STEP.  If 
fibrous  packing  is  used,  the  boxes  should  be  filled  full  and  the  glands 
tightened  down  very  moderately.  The  tightening  of  the  glands  can  best 
be  done  when  steam  is  on  and  the  machine  is  in  motion,  when  they 
should  be  tightened  only  sufficiently  to  stop  leakage  and  no  more.  When 
excessive  tightening  is  required  to  stop  leakage,  the  packing  should 
be  completely  renewed.  Some  pumps  are  fitted  with  metallic  packing. 
This  packing  is  usually  prepared  by  specialists  and  fully  guaranteed. 
Their  directions  for  use  should  be  carefully  followed.  In  case  of  failure  or 
unsatisfactory  results,  the  makers  should  be  consulted. 

THE  OILING  OF  THE  MACHINERY  Is  THE  NEXT  STEP  and  is  a  very 
important  one.  All  rubbing  surfaces  should  be  provided  with  suitable 
oiling  devices  designed  for  the  particular  place  and  service.  The  quality 
of  oil  should  be  carefully  selected  to  suit  the  velocity  and  pressure  of  the 
rubbing  surfaces  on  which  it  is  used.  For  use  within  the  steam  cylinder, 
heavy  mineral  oil  is  the  only  oil  capable  of  withstanding  the  high  temper- 
ature. When  starting  up  new  pumps,  only  the  best-quality  oil  should  be 
employed,  regardless  of  price.  A  liberal  use  of  this  oil  for  the  first  month 
will  go  far  toward  reducing  subsequent  oil  bills. 

A  PUMP  MUST  OFTEN  RUN  CONTINUOUSLY  WITHOUT  INTERRUPTION 
— FOR  A  MONTH  OR  EVEN  LONGER.  This  requires  that  all  oiling  devices 
be  so  arranged  that  they  can  be  replenished  and  adjusted  while  the 
machine  is  in  motion.  It  is  a  good  plan  to  provide  two  sets  of  oiling 
systems  for  all  of  the  principal  journals.  Then,  if  one  fails  the  other  can 
be  used  while  the  disabled  one  is  being  overhauled.  All  oil  holes  are 
generally  stopped  with  wooden  plugs  or  bits  of  waste  twisted  into  the 
hole,  or  are  otherwise  protected  while  the  machine  is  being  erected. 
These  should  now  be  removed  and  the  holes  and  oil  channels  thoroughly 
cleaned.  Bearings  should  be  flooded  with  oil  at  first  to  wash  out  any 
dust  or  grit  that  may  have  reached  the  rubbing  surfaces. 


70  STEAM  POWER  PLANT  AUXILIARIES  [Div.  2 

THE  STEAM  END  Is  Now  READY  To  BE  WARMED  UP.  (From  now  on 
the  method  of  starting  a  pump  is  the  same  whether  the  pump  is  a  new 
or  an  old  one.)  To  warm  up  the  steam  end,  the  throttle  is  opened 
slightly  and,  with  the  drain  cocks  opened  wide,  steam  is  allowed  to  blow 
through  the  cylinder  until  no  more  water  passes  from  the  drain  cocks. 
The  steam  by-pass  pipes  should  be  used  where  multiple-expansion 
pumps  are  being  started.  If  the  pump  has  a  valve  gear  that  can  be 
operated  by  hand,  the  warming  up  can  be  hastened  by  working  the  valve 
back  and  forth  slowly.  While  the  steam  end  is  warming  up,  the  water 
end  should  be  made  ready  by  opening  the  stop-valve  in  the  delivery  pipe 
and  otherwise  insuring  that  the  pump  has  a  free  delivery.  If  a  stop- 
valve  is  fitted  to  the  suction  pipe,  this  should  be  opened.  If  the  pump  is 
compound  or  triple  expansion,  the  water  by-pass  valves  must  be  opened 
until  the  machine  has  made  a  sufficient  number  of  strokes  to  bring  the 
intermediate  and  low-pressure  cylinders  into  action.  Then  the  by-pass 
valves  should  be  closed.  If  the  pump  is  fitted  with  dash-relief  valves, 
these  should  be  closed  before  starting,  keep  the  pistons  as  far  from  the 
heads  as  possible  in  starting.  Should  the  pump  exhaust  into  an  inde- 
pendent condenser,  this  should  be  started  and  a  vacuum  obtained  be- 
fore starting  the  pumps. 

To  START  THE  PUMP,  the  foregoing  precautions  having  been  observed, 
open  the  throttle  slowly.  Permit  the  pistons  to  work  back  and  forth  very 
slowly  a  few  times,  gradually  increasing  the  velocity  until  full  speed  is 
attained.  After  the  pump  has  been  running  a  few  minutes,  close  the 
drain  cocks.  If  the  pump  has  dash-relief  valves,  the  length  of  stroke 
may  now  be  carefully  adjusted. 

To  STOP  THE  PUMP,  close  the  throttle,  open  the  drain  cocks,  and  (if 
there  is  one)  close  the  gate  valve  in  the  discharge  pipe.  Finally  shut 
down  the  condenser.  If  the  pump  is  to  remain  stopped  for  some  time, 
close  the  suction  valve. 


79.  The  Causes  Of  Scoring  Of  Pump-Valve  Stems  and 
Piston  Rods  may  be  one,  or  all,  of  the  three  following:  (l) 
Use  of  an  improper  packing,  as  a  packing  consisting  of  plain, 
unlubricated,  hemp  or  rope  fiber.  (2)  Permitting  a  fibrous 
packing  to  remain  in  the  stuffing-box  after  it  has  become  hard 
and  brittle  through  age.  When  the  packing  attains  this  con- 
dition, attempts  to  prevent  the  steam  from  blowing  out  around 
the  rod  by  drawing  up  on  the  gland  will  inevitably  result  in 
cutting  and  scoring  the  rod.  (3)  Use  of  an  improper  cylinder 
lubricant,  as  an  oil  containing  an  excess  of  animal  fats.  Such 
oils,  in  the  presence  of  high  temperature,  evolve  an  acid  which 
is  particularly  damaging  to  iron  and  steel. 


SEC.  80] 


DIRECT- ACT  I  NO  STEAM  PUMPS 


71 


80.  Table  Showing  Duty  And  Steam  Consumption  Of 
Direct-Acting  Pumps.  Simple,  Non-Condensing  Steam 
Cylinder.  (Values  correct  only  for  the  typical  efficiencies 
which  are  given.  For  other  efficiencies  modify  values 
proportionately.) 


Non-jacketed,  but  lagged;  wire  drawing  =  4.7  lb.;  back  press.  =  16  Ib.  per  sq.  in. 


Boiler  pressure  

50 

70 

90 

100 

110 

120 

150 

Absolute  initial  press  

60 

80 

100 

110 

120 

130 

160 

44 

64 

84 

94 

104 

114 

144 

Card  duty,  million  ft  -lb 

45 

50  5 

53  5 

55 

56 

57 

58  5 

Stroke, 
in. 

Mech. 
effic  , 
per 
cent. 

Steam 
effic., 
per 
cent. 

Total 
effic., 
per 
cent. 

Actual  duty,   million  ft.-lb.  per   1,000  lb. 
dry   steam  =  upper   fig.     Lb.   dry   steam 
used  per  water  h.p.  per  hr.  =  lower  fig. 

9.5 

10.6 

11.3 

11.6 

11.8 

12.0 

12.3 

4 

55.0 

37.5 

21 

208 

187 

175 

171 

168 

165 

161 

11.7 

13.1 

13.9 

14.3 

14.6 

14.8 

15.2 

6 

65.0 

40.0 

26 

169 

151 

143 

139 

136 

134 

130 

13.5 

15.2 

16.1 

16.5 

16.8 

17.1 

17.6 

8 

70.0 

42.5 

30 

147 

130 

123 

120 

118 

116 

113 

15.5 

17.2 

18.2 

18.7 

19.1 

19.4 

19.9 

10 

75.0 

45.0 

34 

128 

115 

109 

106 

104 

102 

100 

16.6 

18.7 

19.8 

20.4 

20.7 

21.0 

21.7 

12 

77.5 

47.5 

37 

119 

106 

100 

97 

96 

94 

91 

18.0 

20.2 

21.5 

22.0 

22.5 

23.0 

23.5 

15 

80.0 

50.0 

40 

110 

98 

92 

90 

88 

86 

84 

19.4 

21.7 

23.0 

23.7 

24.0 

24.5 

25.2 

18 

82.5 

52.5 

43 

102 

91 

85 

84 

83 

81 

79 

21.0 

23.7 

25.1 

25.9 

26.5 

26.9 

27.5 

24 

85.0 

55.0 

47 

94 

83.5 

79 

76 

75 

74 

72 

72' 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  2 


81.  Table  Showing  Duty  And  Steam  Consumption  Of 
Pumps.  Compound,  Non-condensing  Steam  Cylinder  (See 
limitations  in  Table  80.) 


Non-jacketed,  but  lagged;  wire  drawing  =  4.7  lb.;  back  press.   =  16  Ib.  per  sq.  in. 


Boiler  pressure  

50 

70 

90 

100 

110 

120 

150 

Absolute  initial  press  

60 

80 

100 

110 

120 

130 

160 

Ratio  of  cylinders  

1.94 

2.24 

2.5 

2.62 

2.74 

2.85 

3.16 

m.e.p.  on  area  of  h.p.  cyl  

58.0 

88.5 

120.0 

136.0 

152.0 

168.8 

218.8 

Card  duty,  million  ft.-lb  

60.0 

69.5 

76.5 

79.5 

82.0 

84.0 

89.0 

Mech. 

Steam 

Total 

Actual    duty,    million    ft.-lb.    per    1,000    lb.    dry 

Stroke, 

effic., 

effic., 

effic., 

steam  =  upper    fig.     Lb.    dry    steam    per    water 

in. 

per 

per 

per 

h.p.  per  hr.  =  lower  fig. 

cent. 

cent. 

cent. 

15.6 

18.1 

19.9 

20.7 

21.4 

21.8 

23.1 

6 

65.0 

40.0 

26 

127 

110 

99 

95 

92 

91 

85 

18.0 

20.8 

22.9 

23.8 

24.6 

25.2 

25.7 

8 

70.0 

42.0 

30 

110 

95 

86 

83 

80 

78 

74 

20.4 

23.6 

26.0 

27.0 

27.9 

28.5 

30.3 

10 

75.0 

45.0 

34 

97 

84 

76 

74 

71 

69 

65 

22.2 

25.7 

28.3 

29.4 

30.4 

31.1 

33.0 

12 

77.5 

47.5 

37 

89 

77 

70 

67 

65 

64 

60 

24.0 

27.8 

30.6 

31.8 

32.8 

33.6 

35.6 

15 

80.0 

50.0 

40 

83 

71 

65 

62 

60 

59 

56 

25.8 

29.9 

32.9 

34.2 

35.3 

36.1 

38.3 

18 

82.5 

52.5 

43 

77 

66 

60 

58 

56 

55 

52 

28.2 

32.6 

36.0 

37.4 

38.4 

39.5 

41.9 

24 

85.0 

55.0 

47 

70 

61 

55 

53 

52 

50 

48 

30.0 

34.0 

38.2 

39.7 

41.0 

42.0 

44.5 

36 

87.5 

57.5 

50 

66 

58 

52 

50 

48 

47 

45 

82.  Table  Showing  Duty  And  Steam  Consumption  Of 
Pumps.  Compound,  Condensing,  Steam  Cylinder.  (See 
limitations  in  Table  80.) 


SEC.  82] 


DIRECT-ACTING  STEAM  PUMPS 


73 


L-p.  cyl.  jacketed  and  lagged;  wire  drawing  =  4.7  lb.;  back  press.  =  6  Ib.  per  sq.  in. 


70 

90 

100 

120 

150 

170 

180 

80 

100 

110 

130 

160 

180 

190 

3  65 

116.2 

151 

168.5 

203.5 

256 

291 

308.  5 

Card  duty,  million  ft.-lb  

91.0 

96.5 

98  ' 

101.5 

104.5 

106.5 

107 

Mech. 

Steam 

Total 

Actual   duty,    million    ft.-lb.    per    1,000    lb. 

Stroke, 

effic., 

effic., 

effic., 

dry  steam  =  upper  fig.     Lb.  dry  steam  per 

in. 

per 

per 

per 

water  h.p.  per  hr.  =  lower  fig. 

cent. 

cent. 

cent. 

37.4 

39.6 

40.2 

41.6 

42.9 

43.7 

44.0 

10 

75.0 

55.0 

41 

53 

50 

49 

48 

46 

45 

45 

41.0 

43.4 

44.1 

45.5 

47.1 

48.0 

48.0 

12 

77.5 

57.5 

45 

48 

45 

45 

44 

42 

42 

41 

43.7 

46.4 

47.0 

48.7 

50.1 

51.1 

51.3 

15 

80.0 

60.0 

48 

45 

43 

42 

41 

40 

39 

38 

47.4 

50.0 

51.0 

52.8 

54.3 

55.4 

55.8 

18 

82.5 

62.5 

52 

42 

40 

39 

37 

37 

36 

35 

50.0 

53.0 

54.0 

55.9 

57.6 

58.6 

59.0 

24 

85.0 

65.0 

55 

40 

37 

37 

35 

35 

34 

33 

53.8 

57.0 

58.0 

60.0 

61.7 

62.8 

63.1 

36 

87.5 

67.5 

59 

37 

35 

34 

33 

32 

32 

31 

57.2 

60.8 

61.9 

64.0 

65.8 

67.1 

67.4 

48 

90.0 

70.0 

63 

35 

33 

32 

31 

30 

30 

29 

QUESTIONS  ON  DIVISION  2 


1.  What  is  a  double-acting  suction  pump? 

2.  Explain  the  operation  of  a  double-acting  suction  pump. 

3.  What  velocity  of  water-flow  is  recommended  for  the  suction-piping  of  steam 
pumps?     For  the  discharge-pipe  of  a  simplex  pump?     For  the  discharge-pipe  of  a 
duplex  pump? 

4.  What  is  a  piston-pump?     A  plunger-pump? 

5.  What  is  an  outside-packed  plunger-pump?     An  inside-packed  plunger-pump? 

6.  What  is  the  distinction  between  an  outside-end-packed  plunger  pump  and  an 
outside-center-packed  plunger  pump? 

7.  For  what  maximum  discharge-pressures  are  piston  and  plunger  pumps  respectively 
adapted? 

8.  Explain  the  method  of  packing  a  water-piston  with  hydraulic  packing. 

9.  In  what  class  of  pump  service  are  soft  rubber-composition  valve  discs  especially 
suitable?     In  what  class  of  pump  service  are  metal  valve  discs  especially  required? 

10.  How  are  the  water-valves  arranged,  with  reference  to  the  pistons  or  plungers,  in 
horizontal    direct-acting    steam-pumps?     Which    arrangement    is    commonly    used    in 


74  STEAM  POWER  PLANT  AUXILIARIES  [Div.  2 

pumps  for  high-pressure  service?     Which  arrangement  is  recommended  for  vacuum 
pumps? 

11.  What  is  the  function  of  an  air-chamber? 

12.  Explain  the  operation  an  of  air-chamber. 

13.  Why  are  air  chambers  less  necessary  on  duplex  pumps  than  on  simplex  pumps? 

14.  What  is  the  highest  level,  consistent  with  good  service,  to  which  the  water  may 
rise  in  an  air  chamber? 

15.  What  is  a  snifter,  as- used  in  air-chamber  service?     How  does  it  work? 

16.  Describe  a  method  of  recharging  air-chambers  in  pumping  systems  working  under 
pressures  up  about  1,000  Ib.  per  sq.  in. 

17.  What  is  the  proper  ratio  of  air-chamber  volume  to  water-piston  displacement  in  a 
single  pump?     In  a  duplex  pump?     In  fire-pumps? 

18.  What  is  the  function  of  a  vacuum  chamber? 

19.  Explain  the  operation  of  a  vacuum  chamber. 

20.  What  is  a  simplex  steam-pump?     A  duplex  steam-pump? 

21.  What  is  meant  by  the  term  steam-thrown,  as  applied  to  the  steam-valves  of 
simplex  pumps? 

22.  Upon  what  adjustment  does  the  length  of  stroke  of  simplex  pumps  with  external 
valve  gears  commonly  depend? 

23.  Assuming  that  the  crossheads  are  properly  secured  to  the  piston-rods,  what  three 
principal  adjustments  are  necessary  for  correctly  setting  the  steam-valves  of  a  direct- 
acting  duplex  pump? 

24.  What  is  the  three-fold  function  of  the  valve-stem  lost-motion  in  duplex  direct- 
acting  steam  pumps? 

25.  Describe  the  cycle  of  steam-valve  motion  in  the  operation  of  a  duplex  pump. 

26.  What  disadvantage  ordinarily  results  from  incorrect  adjustment  of  the  valve- 
stem  lost-motion  in  duplex  pumps? 

27.  Describe  a  method  of  marking  the  striking-points  of  duplex-pump  pistons. 

28.  How  much  lost-motion  should  the  steam  valve-stems  of  duplex  pumps  ordinarily 
have? 

29.  What  is  meant  by  compression-space  in  the  steam-cylinders  of  duplex  pumps? 

30.  What  are  cushion-valves  on  duplex  pumps? 

31.  What  are  the  advantages  of  simplex  steam-pumps  as  compared  with  duplex 
steam-pumps?     What  are  the  disadvantages?     Which  type  is  recommended  for  fire- 
protection  service  in  buildings?     Which  type  is  recommended  for  use  in  connection 
with  surface  condensers?     Why? 

32.  What  considerations  govern  the  proportioning  of  water-piston  areas  to  steam- 
piston  areas  in  boiler  feed  pumps?     What  is  the  usual  proportion? 

33.  What  two  principal  considerations  govern  the  selection  of  a  direct  steam  driven 
boiler  feed  pump? 

34.  What  are  the  causes  which  may  impair  the  effectiveness  of  a  pump  when  it  is  in 
service? 

35.  Explain  some  conditions  which  may  cause  a  pump  to  fail  to  raise  water.     Give 
remedies  for  each. 

36.  Explain  the  method  of  repairing  the  steam    valve    and    valve   seat   in    a   pump 
when  they  are  badly  worn. 

37.  Enumerate  and  explain  the  successive  steps  in  erecting  a  pump. 

38.  Discuss  steam-pump  lubrication. 

39.  Explain  how  a  pump  should  be  started. 

40.  What  are  the  steps  in  stopping  a  pump? 

PROBLEMS  ON  DIVISION  2 

1.  A  direct-acting  steam-pump  for  low-speed  service  has  a  plunger  diameter  of  12  in. 
The  plunger  is  inside-packed.  The  plunger-rod  is  of  3-in.  diameter.  How  many 
flat  disc  valves,  each  of  4-in.  diameter  and  0.25-in.  lift,  should  there  be  in  each  set  of 
suction  and  delivery  valves  in  this  pump? 


DIVISION  3 
CRANK-ACTION  PUMPS 

83.  Crank-Action  Pumps  include  piston  or  plunger  pumps 
of  all  forms  which  depend  for  their  operation  on  the  circular 
motion  of  a  crank-shaft.  They  may  be  classified  as  follows: 
(1)  Crank-and- fly -wheel  pumps  in  which  the  reciprocating  move- 
ment of  the  pump  piston  or  plunger  is  derived  directly  (Figs. 


frir  Chamber- 

/-Steam  Cylinder 

Vacuum  Chamber 
Steam  Mounted  At 

5upp/y\         Suction  Inlet- . 


,.-Dischotrge  Noziles 


Eccentric-' 
FIG.  84. — Steam-Driven  Crank-And-Fly-Wheel  Pump. 

84  and  85)  or  indirectly  (Fig.  86)  from  the  reciprocating 
movement  of  a  piston  in  a  steam  cylinder  but  is  dependent 
for  its  continuance  upon  the  inertia  effect  of  the  rotative  move- 
ment of  a  crank-shaft  and  fly-wheel.  (2)  Crank-action  power 
pumps  in  which  the  reciprocating  movement  of  the  pump 
piston  or  plunger  is  derived  from  the  rotative  movement  of  a 
mechanically-driven  crank-shaft.  Figures  87,  88  and  89  show 

75 


76 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  3 


belt  driven  power  pump  and  Fig.  90  shows  a  gear  driven  power 
pump. 


liii 


Slide  Yotlw-. 


Suction  Ink  t s-''    Water  Cylinder 

fccentrk  Rod--  , ' — ' 

Steam  Cylinders-     SI 


FIG.  85.— Horizontal    Section    Of    A    Double-Acting    Duplex    Crank-And-Fly-Wheel 

Pump. 


Connecting 
Rod.     9 


Fia.  86. — Crank-Action  Pump  Of  The  Walking-Beam  Type  (Pumps  Of  This  Type  Are 
Now  Practically  Obsolete). 

84.  In  The  Operation  Of  Crank-And-Fly-Wheel  Pumps 
the  steam  is  worked  expansively  in  the  driving  cylinders 
instead  of  being  admitted  during  the  entire  stroke  of  the  piston 
(Sec.  90),  as  in  the  operation  of  direct-acting  steam-pumps. 


SEC.  84] 


CRANK-ACTION  PUMPS 


77 


Hence,  a  fly-wheel  is  necessary  to  insure  approximately  uni- 
form movement  throughout  the  stroke.  (See  the  author's 
STEAM  ENGINES.)  The  pump  piston  or  plunger  is  usually 
connected  directly  to  the  piston-rod  (Fig.  91)  of  the  driving 
cylinder.  Hence,  the  function  of  the  crank-and-fly-wheel  is 
only  to  insure  minimum  variation  of  the  rotative  speed. 


Tight 
Pulley 


Loose  Pulley 


Tight  Pulley. 


•D/'sc- . 
Crank 


FIQ.  87. — A   Belt-Driven  Single-Acting 
Pump  For  Boiler-Feeding. 


FIG.  88. — Combination    High-Service    And 
Low-Service  Belt-Driven  Pumps. 


Crank-and-fly-wheel  pumps  are  generally  more  economical 
than  direct-acting  steam  pumps.  This  is  due  to  the  expansive 
use  of  steam  in  the  cylinders  and  to  the  better  valve  action 
which  is  obtained,  as  in  the  steam  engine,  by  the  use  of 
properly-designed  Corliss  and  slide  valve-gears.  Hence,  they 
are  chiefly  employed  where  steam-driven  pumps  are  desired 
but  considerations  of  economy  preclude  the  application  of  the 
direct-acting  type. 


78 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


NOTE. — AN  ADVANTAGE  CLAIMED  FOR  CRANK  ACTION  As  COMPARED 
WITH  DIRECT  ACTION  in  the  operation  of  steam-pumps  is  that  crank- 


T/yM  Pulley..., 
•Disc  Crank 


\ 


~ Discharge  Valves—""" 


Fia.  89. — A  Belt-Driven  Double-Acting     FIG.  90. — Belt  Driven  Single-Acting  Plunger 
Pump  For  Boiler  Feeding.  Pump  For  Boiler  Feeding. 


.flywheel 


Suction  • 
Inlet----'' 

Fia.  91. — Single-Acting  Crank- And-Fly wheel  Pump  For  Hydraulic  Elevator   Service. 

action  entirely  obviates  the  short-stroking  of  the  pistons  (Sec.  75)  which 
is  liable  to  occur  with  direct-acting  pumps.     Also,  since  the  limits  of 


SEC.  85] 


CRANK-ACTION  PUMPS 


79 


the  piston  stroke  are  definitely  fixed,  less  clearance  is  necessary  at  the 
ends  of  the  cylinders.  Crank  action,  as  a  rule,  permits  of  a  higher  piston 
speed  than  is  practicable  with  direct-action.  This  IP  due  to  the  energy 
which  is  stored  up  in  the  moving  mass  of  the  fly-wheel  at  the  termination 
of  the  stroke.  This  energy  is  available  for  reversing  the  motion  of  the 
piston.  With  direct-action,  the  reversal  of  the  stroke  is  effected  solely 
by  steam  pressure. 

NOTE. — STEAM-DRIVEN  PUMPS  OF  THE  CRANK- AND-FLY- WHEEL  TYPE 
WERE  FORMERLY  EXTENSIVELY  USED  IN  CITY  WATER  WORKS  and  large 
hydraulic  elevator  installations  (Fig.  91).  The  comparatively  large  units 
designed  for  this  class  of  service  are  called  pumping  engines.  Pumps  (Fig. 
84)  which  are  used  in  sugar  mills  for  pumping  molasses  are  also  of  this 
type. 


Air  Eccenfr/c--. 


-  ^  ,       -  i 

rVertral  Section  Of    ff-Cross-secfion  Of  \ 
m 


FIG.  92. — Alberger  Rotative-Reciprocating  Dry- Vacuum  Pump. 

NOTE. — A  MAJORITY  OF  STEAM  AIR-COMPRESSORS  AND  DRY- VACUUM 
PUMPS  are,  strictly  speaking,  included  in  this  group  but  are  more  con- 
veniently discussed  under  other  headings.  For  vacuum  pumps  (Fig. 
92)  see  Sec.  354. 

85.  The  Steam  Consumptions  Of  Crank-And-Fly-Wheel 
Pumps  are  determined  by  the  same  general  factors  that  govern 
the  steam  consumptions  of  steam  engines  in  similar  classes  of 
service.  These  factors  are,  mainly,  the  type  of  steam  valve 
gear  that  is  used  and  the  methods  of  operation — whether  sim- 
ple or  compound,  condensing  or  non-condensing.  Slow-speed 
crank-and-fly-wheel  pumps  with  single  steam  cylinders  of  the 
simple  slide-valve  type  consume,  when  operated  non-conden- 


80 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


sing,  about  50  Ib.  of  steam  per  indicated  horse  power  hour. 
High-duty  crank-and-fly-wheel  pumps  with  compound  steam 
cylinders  and  Corliss  steam  valves  consume,  when  operated 
non-condensing,  about  25  Ib.  of  steam  per  indicated  horse 
power  hour.  With  condensing  operation,  the  steam  consump- 
tion of  these  high-duty  pumps  may  be  as  low  as  10  Ib.  of  steam 
per  indicated  horse  power  hour. 

86.  The  Advantages  And  Disadvantages  Of  Crank-And- 
Fly-Wheel  Pumps  in  comparison  with  direct-acting  steam- 
pumps  may  be  enumerated  as  follows:  (1)  Steam-consumption 


Cro/nff  Shaft-. 


Main  Gear- 


/ Loose  Pulley 
!    fTfyht  Pulley 


\  ."Connecting -Rods 


Plungers 


Discharge  Outlet- 

FIG.  93. — Triplex  Pump  For  Heavy  Liquids. 


is  generally  more  economical.  (2)  May  be  run  at  higher  speeds 
for  most  classes  of  service.  (3)  First  cost  is  greater.  (4) 
Require  greater  operating  attendance.  (5)  Cost  of  maintenance 
is  greater. 

NOTE. — The  water-ends  of  crank-action  pumps  are  built  in  many 
respects  like  the  water-ends  of  direct-acting  pumps,  which  are  discussed 
in  the  preceding  Div.  The  information  there  given  relative  to  the  care 
of  valves,  packing,  and  management  in  general  applies  here  to  pistons, 
glands,  plungers,  and  other  parts.  The  subjects  of  piping,  pressures, 
heads,  suction  ^ind  the  like  are  also  largely  omitted  here  as  they  are  dis- 
cussed in  Divs.  1  and  2. 


SEC.  86] 


CRANK-ACTION  PUMPS 

Main  focrr— ._ 


81 


Pinion-.^ 


''-Discharge  Valves'  '  *  -Suction  Inlet 
FIG.  94.— Sectional  View  Of  Single-Acting  Triplex  Pump. 


Crank 
Shaft—.. 


Main  GeaK, 
A' 


V. 

Guides-.. 

.Cross neofot  Pin-. 
Discharge  Valve, 


FIG.  95. — Sectional  View  Of  Double-Acting  Triplex  Pump. 


82 


STEAM  POWER  PLANT  AUXILIARIES 


IDiv.  3 


87.  Crank-Action  Power  Pumps  may  be  divided  into  three 
main  classes:  (1)  Simplex  power  pumps  (Fig.  87)  in  which  the 
pumping  operation  is  performed  by  a  single  piston  or  plunger, 


Spur  Gear  On  Crynk  Shaft, 
Crank-. 


FIG.  96. — Pump  Driven  By  Motor  Through  Spur  Gearing. 
.-••Crunk  Shaft  Pulley 


FIG.  97. — A  Belt-Driven  Power-Pump. 


Crank  Shaft 
(Sprocket  Wheel 


FIG.  98. — A  Chain-Driven  Power  Pump. 


(2)  Duplex  power  pumps  (Fig.  90,  200,  and  201)  in  which  the 
pumping  operation  is  performed  by  two  pistons  or  plungers 
operated  by  a  common  crank-shaft.  (3)  Triplex  power  pumps 


SEC.  87] 


CRANK-ACTION  PUMPS 


83 


(Fig.  93)  in  which  the  pumping  operation  is  performed  by 
three  pistons  or  plungers  operated  by  a  common  crank-shaft. 
These  pumps  may  all  be  single  acting  (Fig.  94)  or  double 
acting  (Fig.  95).  If  the  pump  is  double  acting,  the  plunger 
may  be  in  two  parts  as  in  Fig.  53. 


Pump- 


i 


FIG.  99. — "Goulds"  Triplex  Deep  Open- Well  Pump. 


NOTE. — Power  may  be  supplied  to  power  pumps  by  electric  motors 
(Fig.  96),  gas  or  gasoline  engines,  water-wheels,  steam  engines  or  line-shaft- 
ing variously  driven.  This  power  may  be  transmitted  to  the  crank-shafts 
of  the  pumps  by  means  of  belts  (Fig.  97),  chains  (Fig.  98),  gears  or  rope- 
drives.  The  pump  crank-shafts  may  also  be  connected  directly  to  the 
drive-shafts  of  the  prime  movers. 


84 


STEAM  POWER  PLANT  AUXILIARIES 


[Dw.  3 


88.  Crank-Action  Power  Pumps  Are  Designed  And 
Arranged  In  Various  Ways  For  Deep-Well  Service. — Since 
wells  are  frequently  more  than  22  feet  (practical  suction  lift, 
Sec.  2)  deep,  it  is  often  necessary  to  install  pumps  with  their 
cylinders  below  the  ground  level  so  as  to  force  the  water  out 


Wafer -Level  In  Tank-.. 


Fio.  100. — A  Motor-Driven  Deep- Well  Pump. 

by  pressure.  Sometimes  wells  have  large  sectional  areas  and 
are  comparatively  shallow.  For  such,  the  under-ground  por- 
tions of  the  pumps  may  be  installed  (Fig.  99)  very  much  like 
ordinary  power  pumps.  They  are,  however,  provided  with 
elongated  plunger-rods  which  connect  to  the  crank-shafts 


SEC.  89] 


CRANK-ACTION  PUMPS 


85 


Plunger 


located  above  ground.  More  often  deep  wells  are  merely 
drilled  holes  ranging  possibly  from  2  inches  to 
12  inches  in  diameter  protected  by  metal-tube 
casings.  They  may  be  several  hundred  feet 
deep.  For  such  wells  it  is  necessary  to  use 
the  so-called  deep-well  or  artesian-well  pumps 
(Fig.  100)  which  have  been  especially  de- 
signed for  this  service. 

89.  Crank-Action  Pumps  For  Deep-Well 
Service  are  of  three  kinds :  ( 1 )  Single  -acting 
pumps  discharging  on  the  up-stroke  only 
(For  exception  see  Sec.  90),  Fig.  101.  (2) 


Driving-Pool ., 
Connection--' 

Brass  ball-Valve- 
For  Down-Stroke 


Brass  Cylinder 
Shell 


Rubber  Disc-Valve 
For  Down-Stroke  / 
Suction -' 


Ducts  Connecting 
Outer  And  Inner 
Plunger-Tubes-.. 


Rubber  Disc-Valve- -- 
For  Up-5troke  Suction^ 


-Drop  Pipe 

Rubber  Disc-Valve 
For  Up-Stroke 
Discharge*. 


•Flange  Leather 
Packing 


Cup- Leather  Plun- 
ger Packing* 


Flange,  Leather 
Packing* 


""Hollow  Tail-Roof 


FIG.  101.— Cylinder 
Of  Single-Acting  Deep- 
Well  Pump. 


FIG.  102.— Cylinder  Of  Double-Acting  Deep- Well 
Pump  (Plunger  Making  An  Upward  Stroke). 


Double-acting  pumps  having  one  plunger  but  discharging  on 
both  the  up-stroke  and  the  down-stroke,  Fig.  102.     (3)  Two- 


86 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


stroke  pumps  having  two  plungers 
operating  in  one  cylinder  controlled 
by  two  well-rods,  Figs.  103  and  104. 
Pumps  of  this  last  type  discharge 
almost  continuously  and  are  fre- 
quently used  in  deep- well  service. 

NOTE. — SOME  ENGINEERS  PREFER  AN 
AIR-LIFT  for  certain  deep-well  pumping 
applications,  because  it  has  no  moving 
members  (except  the  compressor),  is  inex- 
pensive, and  has  no  parts  requiring  repair 
underground  where  they  are  inaccessible. 
They  are  not  as  efficient  from  a  power 


sCrotnkGeoirs 


Pulley 


Gurofe  Roofs 


Discharge  Pipe 


FiQ,     103. — Chippewa    Power-Driven    Deep-Well- 
Pump  Head  Or  Operating  Gear. 


FIG.  104. — Deep-Well  Pump 
Cylinder  Fitted  With  Differ- 
entially-Operating Plungers. 


SEC.  90] 


CRANK-ACTION  PUMPS 


87 


standpoint  as  pumps  but  are  proof  against  damage  by  grit  and  are  not 
likely  to  get  out  of  order. 

EXPLANATION. — Fig.  100  shows  a  typical  motor-driven  deep-well  in- 
stallation. It  may  be  single-acting  if  used  with  the  cylinder  and  plunger 
of  Fig.  101  or  double-acting  if  used  with  the  cylinder  and  plunger  of 
Fig.  102.  The  lower  ball-valve  (Fig.  101)  opens  on  the  up-stroke  allow- 
ing the  pump  to  fill  with  water.  On  the  down-stroke,  the  lower  valve 
seats  and  the  upper  valve  opens  allowing  the  water  in  the  cylinder  to 
flow  past  the  plunger.  On  the  next  up-stroke,  the  water  is  lifted  up  the 
drop-pipe.  The  double-acting  plunger  (Fig.  102)  operates  similarly  to 
the  single-acting  plunger  on  the  up-stroke.  On  the  down-stroke,  how- 
ever, the  water,  instead  of  merely  passing  tl)e  plunger,  is  forced  up  the 
drop-pipe  through  the  hollow  plunger-rod.  Meanwhile  more  water  is 
drawn  into  the  upper  part  of  the  cylinder  through  the  hollow  tail-rod. 

90.  A  Compound  Or  Two-Stroke  Deep-Well  Pump  Operat- 
ing Gear  is  shown  in  Fig.  103.  Its  plungers  and  cylinder, 
which  are  located  underground,  are  sim- 
ilar to  those  shown  in  Fig.  104.  The  two- 
stroke  type  of  pump  has  the  advantage 
over  the  single-acting  type  that  it  insures 
a  more  nearly  continuous  movement  of 
vertical  water  column.  Its  advantage 
over  the  single-plunger  type  is  that  the 
two  plungers  are  of  about  the  same 
weight  and  balance  each  other;  as  one  is 
going  up,  the  other  is  coming  down. 

EXPLANATION. — As  the  geared  cranks  A  and  B 
(Fig.  103)  revolve,  one  or  the  other  of  the  two 
plungers  L  and  T  (Fig.  104)  is  on  the  up-stroke 
continuously,  except  at  dead-center.  When  the 
plunger  T  is  on  the  up-stroke,  its  valve  Vi  seats 
and  water  is  forced  by  it  up  the  drop-pipe,  while 
valve  V2  (Fig.  105)  opens  and  allows  water  to 
pass  plunger  L  which  is  then  on  the  down-stroke. 
On  the  return  stroke,  the  valve  V2  seats  and 
water  is  forced  through  valve  Vi  and  on  up  the 
pipe. 

NOTE. — Most  deep-well  pump  plungers  are  packed  with  leather  cup- 
washers  (Fig.  106).  The  plunger  rods  at  the  top  of  the  drop  pipes  are 
packed,  usually,  with  fibrous  packing  in  the  same  way  as  are  piston-rod 
glands.  The  valves  used  are  either  ball-valves  (Fig.  101),  disk  valves  or 
conical  seated  valves  (Fig.  107).  Plunger-rods  or  well-rods  of  the  single- 
acting  type  are  usually  of  wood  with  steel  fittings  and  should  be  fitted 


''Three  Cup  Washers' 
FIG.  105. — Cross  Section 

^m1PQ4Plunger  Shown 


88 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


with  guide-couplings  (Fig.  108)  which  slide  on  the  inside  of  the  drop- 
pipes  and  prevent  the  rods  from  buckling.  Well-rods  for  double-acting 
pumps  are  generally  made  of  wrought  iron  pipe,  on  account  of  the  com- 


Sleeve- 


Fotge  Chamfered-.^ 


FIG.    106. — Leather  Cup  For  Packing 
A  Deep- Well  Pump  Plunger. 


FIG.   107. — Plunger-Valve  For  Deep- 
Well   Pump. 


pression  strain  on  the  down  stroke.  Guide-couplings  should  be  used 
about  every  twenty  feet.  Two-stroke  pumps  have  a  solid  steel  or  iron 
rod  (Fig.  109)  driving  the  lower  plunger  (Fig.  110).  This  rod  slides 


FIG.  108. — Steel  Guide  Coupling  For  Well-Rods  Of  Deep- Well  Pumps. 

inside  of  a  hollow  tube  or  pipe  which  drives  the  upper  plunger  (Fig.  111). 
Both  rods  must  be  guided  and  packed.  In  open-well  pump  installations, 
the  plunger  rods  are  guided  (Fig.  99)  with  grooved  rollers. 


•  •  -Threaded  Shank  '•-..  section  Of  Solid  Con  nee  ting  -  Pool 

...--•Threaded  End      Section  Of  Hollow  Connecting-Pod-., 


Couplings-, 


FIG.  109.— Connecting  Rods  For  Operating  Plungers  Of  Two-Stroke  Deep- Well  Pump. 


91.  The  Characteristics  Of  Crank-Action  Pumps  are  very 
different  from  those  of  pumps  of  the  direct-acting  type.  Com- 
pare the  indicator  diagrams  for  the  steam-  and  water-ends  of 


SEC.  91] 


CRANK-ACTION  PUMPS 


89 


the  crank-and-fly- wheel  pump  shown  in  Fig.  112  with  cor- 
responding ones  for  direct-acting  pumps  shown  in  Figs.  22 
and  23.  The  difference  between  the  diagram  for  the  steam- 
end  in  Fig.  112  and  that  in  Fig.  22  is  due  to  cut-off  at  about  one 
third  stroke  in  the  crank-action  pump  and  non-expansive 
use  of  steam  in  the  direct-acting  pump.  The  difference 
between  the  water-end  diagram  shown  in  Fig.  112  and  that 
shown  in  Fig.  23  is  due  partly  to  the  more  rapid  movement  of 


Lift  Of 
Valve,-, 


•Suction 
Roc/ 

-Lock- 
Nut 


Fia.  110. — Lower  Plunger  Of  Two-Stroke 
Deep-Well  Pump. 


Fia.  111.— Upper  Plunger  Of  Two-Stroke 
Deep- Well  Pump. 


the  piston  in  mid-stroke  in  the  crank-action  pump  and  uniform 
movement  throughout  the  stroke  in  the  direct-acting  pump. 
The  type  of  water-end  diagram  of  Fig.  112  is  characteristic 
only  for  low-pressure  and  high-speed  crank-and-fly-wheel 
and  power  pumps.  Higher  pressures  and  lower  speeds  in 
crank-action  pumps  produce  indicator  diagrams  which  are 
more  nearly  rectangular.  The  higher-speed  water-end  dia- 
grams are  characterized  by  sharp  pressure  peaks  and  very 
irregular  pressure  curves. 


90 


STEAM  POWER  PLANT  A  UXILI ARIES 


[Div.  3 


Atmospheric  Line-* 
Stroke    Inches    (Reduced) 


FIG.  112. — Steam-End  And  Water-End  Indicator  Diagrams  For  Small  Low- Pressure 
Crank-And-Fly-Wheel  Pump. 


Grotph  Showing  Intermittent  Suction--. 
Graph-  Showing  Intermittent  Discharge--, 


gr 

: 

s, 

— 

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x 

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If         .Z70"          560"         90           B( 
osition   Of  Crank 

FIG.  113. — Graph  Showing  Rates  Of  Suction  And  Discharge  Of  A  Simplex  Single- 

Acting  Pump. 


Line,  Of  No  Suction  And  No  Dfscharge-^ 
Point  Of  No  Sue  tion  And  No  Discharge-. 


•gO  90          ISO          270         360         90  160 

"^Angular   Position      Of    Cranks 

FIG.  114. — Graph  Showing  Rates  Of  Suction  And  Discharge  Of  The  Individual 
Cylinders  Of  A  Duplex  Single-Acting  Pump  With  Cranks  180  deg.  Apart  Or  Of  A 
Simplex  Double-Acting  Pump. 


SEC.  92] 


CRANK-ACTION  PUMPS 


91 


92.  The  Rate  Of  Suction  And  Discharge  Graphs  for  a  sim- 
plex single-acting  crank-action  pump  are  shown  in  Fig.  113. 
The  lines  of  no-discharge  and  no-suction  are  separated  to  show 
the  intermittent  nature  of  the  action  of  pumps  of  this  type. 


fPos/f/bn  Shown  In  Fly.  117  firaph  Of  Total  Discharge  Rctfe  Of  Pump 
30'  \   60'     90"    110'   150*    18$    Z10*     Z4-0*  ITO'    300'  330*   360'    30*     6( 

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Fia.  115. — Graph  Rates  Of  Suction  And  Discharge  Of  The  Individual  Cylinders  (A, 
B  and  C,  Fig.  117)  Of  A  Single-Acting  Triplex  Pump.  Also  The  Resultant  Or  Total 
Discharge  Of  All  Of  The  Cylinders. 

Fig.  114  shows  graphically  the  suction  and  discharge  rates  for 
a  single-acting  duplex  pump  with  cranks  180  deg.  apart.  A 
double-acting  simplex  pump  has  the  same  characteristics. 
Pumps  of  these  types  have  instants  of  inaction  at  the  ends  of 
the  strokes  as  shown  at  A  on  the  graphs. 

EXPLANATION. — THE  SUCTION  AND  DISCHARGE  GRAPHS  FOR  A  TRI- 
PLEX PUMP  similar  to  the  one  shown 
in  Fig.  93  are  shown  in  Fig.  115.  A 
pump  of  this  type  has  a  crank-shaft 
(Fig.  116)  having  three  cranks  which 
are  set  120  deg.  apart.  Fig.  117  shows 
diagrammatically  the  position  of  each 
crank  separately  at  a  given  instant. 
The  graphs  in  Fig.  115  show  the  rates 
of  discharge  and  suction  of  each  of 
the  individual  cylinders  A,  B,  and  C  FlG- 
(Fig.  117)  and  also  of  the  whole  pump. 
The  line  XX'  (Fig.  115)  represents  the  position  of  the  plungers  at  the 
instant  considered  in  Fig.  117.  The  distances  of  the  points  A',  B',  and  C' 
from  the  line  of  zero  suction  and  zero  discharge  (Fig.  115)  represent  the 
rates  at  which  the  cylinders  A,  B,  and  C  are  sucking  or  discharging 
at  the  instant  considered.  Cylinder  A  is  at  dead-center  and,  therefore, 


Spur  Wheel- 
fCroink  Shaft 


''••Cranks--' 


116. — Main    Gear    And    Crank- 
Shaft  Of  A  Triplex  Power  Pump. 


92 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


point  A'  is  on  the  zero  discharge  line.  Cylinder  B  is  nearing  its  maximum 
rate  of  discharge  as  shown  by  the  rise  of  graph  B  at  B'.  Cylinder  C 
has  passed  its  maximum  rate  of  suction,  as  shown  by  the  upward  slope 
of  the  graph  C  at  C'.  Graph  Y  represents  the  total  discharge  rate  and 
graph  Z  the  total  suction  rate  of  the  pump. 


.I20«      ,-Crofnk-.       \2Q'  Directbn  Of  Rotation-, 


Cmnk 
Circle^ 


•£* 


FIG.  117. — Diagrammatic  Illustration  Of  A  Single-Acting  Triplex  Pump,  Showing 
Relative  Positions  Of  Its  Elements  At  A  Given  Instant.  (See  preceding  illustration 
for  graphs.) 

93.  The  Allowable  Speed  For  Crank -Action  Pumps  varies 
over  a  wide  range  according  to  conditions.     The  following 
values  are  from  various  sources: 

94.  Table    Showing    Typical    Crank-Action-Pump    Piston 
Speeds. 


Type  of  pump 


Piston  speed, 

feet 
per  minute 


60  inch  stroke  crank-and-fly-wheel  pump 

30  inch  stroke  crank-and-fly-wheel  pump 

15  inch  stroke  crank-and-fly-wheel  pump 

18  inch  stroke  geared  power  water  pump 

Deep-well  pumps  about  24  inch  stroke 

Water  supply  pumps  5"  X  12"  to  9"  X  16"  50  Ib.  to 

1000  Ib.  pressure 

Water  supply  pumps  5"  X  12"  to  9"  X  16"  up  to 

3000  Ib.  pressure 

Hydraulic  pumps  up  to  5000  Ib.  pressure  


300 
250 
200 
100 
100 

100 

80 
50 


NOTE. — For  thick  liquids  and  high  suction  lifts  the  allowable   piston 
speeds  are  lower  than  specified  above. 


SEC.  95] 


CRANK-ACTION  PUMPS 


93 


95.  Selection  Of  Pumps  For  Liquids  Other  Than   Water 

(MARKS'    HANDBOOK)    should    be  discussed  usually  with  the 
pump  manufacturers.     The  following  indicates  usual  practice: 


Liquid 

Material 

Liquid 

Material 

Brine 
Caustic 
Hydrochloric  acid 

Brass  fitted 
All  iron 
Lead  lined 

Oil 
Sewage 

Brass  fitted 
Brass  fitted 
Large  openings 

NOTE. — CORROSIVE  LIQUIDS  are  handled  ordinarily  by  air  pressure  or 
in  properly-lined  centrifugal  pumps.  Gummy  liquids  are  handled  pref- 
erably in  pumps  with  large  ball-valves.  Volatile  non-corrosive  liquids, 
such  as  alcohol  and  gasoline,  may  be  handled  the  same  as  water  except 
that  the  liquid  must  always  flow  to  the  pump  by  gravity. 

96.  Selection  Of  Proper  Pump  Power  And  Capacity  is  a 

matter  of  computation,  as  explained  in  Div.  1,  but  the  follow- 
ing table  of  typical  pump  data  shows,  in  a  general  way,  the 
size  and  power  necessary  for  a  given  capacity. 

97.  Table  of  Typical  Crank -Action  Pump  Data. 


Power  re- 

Type 

Bore 
and 
stroke, 
inches 

Speed, 
r.p.m. 

quired  in  h.p. 
per  100  Ib. 
per  sq.  in. 
head 

Pulley 
size,  in. 

Capacity, 
gal.  /mill. 

a          d 

2X2 

65 

*0.24 

12  X  \1A 

*3.4 

IV* 

3X4 

55 

*1.00 

14  X  3 

*13 

"go 
<S    K    =?    £ 

4X6 

55 

*3.11 

18  X  3>£ 

.*35 

~  "ail  ^ 

6X8 

50 

*6.40 

20  X  5 

*95 

•S-S  §1 

4  X  12 

42 

*3.96 

18  X  3>^ 

*54 

00      Q 

10  X  12 

40 

*23.00 

24  X  6 

*320 

d 

o 

2X2 

60 

0.32 

12  X  2 

4.7 

3X4 

55 

1.52 

15  X  3 

20 

f  JS 

4X6 

55 

4.65 

20  X  4^ 

53 

"Sr^ 

6X8 

50 

11.10 

30  X  6 

146 

|H 

8  X  10 

45 

21.00 

36  X  6 

292 

*  Duplex  double-acting  pumps  at  the  same  speed  give  approximately 
twice  these  capacities  and  require  twice  the  power  and  pulley  width. 


94 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


NOTE. — CRANK-AND-FLY-WHEEL  PUMP  SIZES  cannot  be  figured  from 
the  relative  boiler  pressure  and  working  pressure  as  can  direct-acting 

pump  sizes  (Sec.  50).  Because  of  the 
cut-off  at  partial  stroke  of  crank-and- 
fly-wheel  pumps,  the  horsepower  of 
the  steam  cylinders  must  be  found 
and  the  capacity  figured  as  for  power 
pumps. 


Cross  head-.. 

Be/i-  Idler. 

Driving 
Motor*. 


98.  The  Advantages  Of  The 
Electrically -Driven  Pumping 
Unit  are :  (1)  It  may  be  located 
many  miles  from  the  source  of 
power  and  still  operate  with  very 
high  efficiency.  These  values  are 
typical: — Line  efficiency,  90  per 
cent.  Motor,  85  per  cent.  Pump 
and  gearing,  82  per  cent.  Over- 
all efficiency  63  per  cent.  For 
steam  or  air-operated  pumps 
which  are  installed  a  consider- 
able distance  from  the  source  of 
power,  the  over-all  efficiency 
would  probably  be  under  25  per 
cent.  (2)  Automatic  control  is 
effected  readily  with  electricity. 
Electrically-driven  pumps  may 
be  started  and  stopped  by  a 
float-operated  switch  which  will 
maintain  a  required  level  in  the 

FIG.  117 A—  Vaile-Kimes   Single-     Supply      tank.       ElectHcally- 

Acting  Deep-well  Pump  Provided  with    Operated  pumps  may  readily  be 

Differential  Piston  For  Securing  Con- 
tinuous Discharge.  (The  plunger,  c,  on    controlled  from  any  reasonable 

its  up  stroke  discharges  half  of  its  dis-     distance 
placement  out  the  discharge,  F,  or  into 


the  air-chamber,  H.  The  other  half  is 
drawn  into  differential  cylinder,  E,  by 
the  upward  movement  of  D.  On  the 
down  stroke,  A  closes  and  the  water  in 
E  is  discharged  out  F  by  D.) 


NOTE. — The  choice  of  a  method  of 
driving  a  boiler  feed  pump  is  dis- 
cussed in  Sees.  214  to  219.  The  prin- 
ciples outlined  therein  are  of  general 

application,  and  are  useful  in  selecting  driving  means  for  a  variety  of 

purposes. 


SEC.  99]  CRANK-ACTION  PUMPS  95 

99.  Simplex  Double -Acting  Pumps  are  manufactured  for  a 
great  variety  of  purposes.     Many  non-corrosive  oils,  solutions 
and  other  liquids  are  handled   in  factories  by  such  pumps. 
The  sizes  range  ordinarily  from  around  2  in.  bore  and  stroke  to 
around   6   in.   bore   and   stroke   for   general   service.     Small 
water-supply  systems  can  often  be  served  effectively  by  pumps 
of  this  simple  type.     Simplex  pumps  of  small  capacity  have 
the  advantages  of  lower  first  cost  and  greater  ease  of  repair 
than    more    complicated    pumps.     In    the    larger    capacities 
these  advantages  disappear.     Simplex  pumps  are  seldom  de- 
signed single-acting  because  of  the  intermittent  discharge  due 
to  such  action.    Single-acting  deep-well  pumps  are  an  exception 
but  the  discharge  is  made  regular  in  some  such  pumps  by  a 
differential  cylinder,  which  is  located  near  the  discharge  out- 
let and  discharges  half  the  water  on  the  upstroke  and  half 
on  the  down-stroke  (Fig.  117 A). 

100.  The  Use  Of  Duplex  Single-Acting  Pumps  is  confined 
largely  to  a  few  special  applications  where  it  is  necessary  to 
reduce  the  first  cost  below  that  of  a  triplex  pump.     They  are 
now  made  seldom,  if  ever.     The  intermittent  discharge  may 
be    a  decided   disadvantage.     For  the  average  service,  the 
duplex  single-acting  pump  has  no  advantage  over  the  standard 
simplex  double-acting  pump. 

101.  Crank-And-Fly-Wheel   Pumps   range   in  size  up   to 
perhaps  10  ft.  stroke  by  4  ft.   bore  for  municipal  pumping 
service.     The  large  pumps  of  this  type  are  usually  compound 
duplex  or  triple  expansion  triplex.     Crank-and-fly-wheel  pumps 
are,  on  account   of  their  high  economies,   used   occasionally 
for  medium  duty,  although  their  first  cost  is  greater  than  that 
of  either  the  centrifugal  or  direct-acting  pumps  with  which 
they  are  in  competition. 

NOTE. — Centrifugal  pumps  driven  by  motors  or  steam  turbines  are 
superseding  crank-and-fly-wheel  pumps  for  large  municipal  pumping 
installations.  The  centrifugal  unit  usually  deteriorates  less  in  efficiency 
with  constant  use  than  does  the  reciprocating  unit.  Furthermore,  the 
much  smaller  size  and  weight  of  the  centrifugal  unit  for  a  given  capacity 
make  its  installation  less  expensive.  These  features  are  conducive  to 
lower  annual  costs. 


96  STEAM  POWER  PLANT  AUXILIARIES  [Div.  3 

102.  Duplex  Double -Acting  Power  Pumps  are  manufactured 
in  sizes  ranging  from  perhaps  2  in.  bore,  4  in.  stroke  to  14  in. 
bore,  12  in.  stroke  for  mine  pumping,  boiler  feeding  (in  the 
smaller  sizes),  drainage  and  general  water-supply  purposes. 
The  additional  parts  necessary  for  the  two  cylinders  of  these 
pumps  are  justified  by  the  smaller  size  of  the  parts  and  the 
better   characteristics  of  the  duplex  pump.     The  cranks  of 
these  pumps  are  usually  set  90  deg.  apart  so  as  to  give  four 
maximum  discharge  peaks  per  revolution. 

103.  Triplex  Single -Acting  Power  Pumps  are  in  competition 
with  duplex  double-acting  power  pumps  for  most  classes  of 
service  and  the  choice  of  design  varies  with  the  manufacturer. 
The  triplex  single-acting  is  a  more  compact  upright  type  of 
pump.     The  duplex  double-acting  type  is  more  common  in  the 
horizontal  design  because  of  the  extra  length  of  guides  neces- 
sary for  double  action.     There  is  some  advantage  in  the  triplex 
single-acting  construction  for  hydraulic  press  work  because  the 
strains  are  more  easily  taken  care  of  by  the  single-acting  form 
of   plunger  and   connecting  rod.     Triplex  pumps  are  more 
commonly  used  than  are  duplex  pumps. 

104.  Triplex  Double-Acting  Pumps  are  used  occasionally 
for  certain  special  applications  in  large  units  for  high-pressure 
pumping.     For  the  average  application  they  possess  no  advan- 
tage over  single-acting  triplex  pumps.     There  are  compara- 
tively few  in  use. 

NOTE. — Multi-stage  centrifugal  pumps  are  now  used  for  many  ser- 
vices where  it  was  formerly  considered  that  the  head  or  pressure  was  too 
high  for  a  centrifugal  pump  to  work  against.  The  efficiency  of  a  cen- 
trifugal pump  is  usually  somewhat  less  than  that  of  a  new  crank-action 
pump.  However,  the  centrifugal  pump  has  advantages  such  as  compact- 
ness, simplicity,  low  up-keep  and  long-continued  efficiency  that  under 
many  conditions  offset  this  disadvantage. 

105.  The  One  General  Rule  In  Selecting  A  Pump  is  first 
to  find  which  types  of  pumps  will  satisfy  the  capacity  require- 
ments of  the  service  being  considered  and  be  reliable  under 
the  conditions.     In  so  doing,  consider:  (1)  Liquid  to  be  handled. 
(2)  Attention  required.     (3)  Characteristics.     (4)  Capacity,  head 
and  power.     Then,  the  eligible  types  having  been  determined, 
select  that  type  which  will  show  the  least  annual  cost  or  the 


SEC.  106] 


CRANK-ACTION  PUMPS 


97 


least  cost  for  pumping  a  certain  quantity  of  water,  on  the 
basis  of:  (1)  Interest  on  investment.  (2)  Depreciation.  (3) 
Maintenance.  (4)  Power  cost.  Often  this  determination 
may  be  made  most  conveniently  on  the  basis  of  pumping  the 
quantity  of  liquid  which  the  pump  must  handle  in  a  year. 
106.  Modern  Pump  Applications. — The  words  in  the 
spaces  (Fig.  118)  refer  only  to  crank-action  pumps.  It  is 
understood  that  only  one  pump  at  a  time  is  being  considered. 
Greater  capacities  can  be  obtained,  of  course,  by  installing 
several  pumps  in  parallel.  Greater  heads  can  sometimes  be 


(Jsual  Limit  Of 
•D/recf-Acfrna  Steam 
Pump  Appl/cctffon 


-Usual  Limit  Of  Centrifugal 
Pump  Application 


FIG.   118. — Modern  Pump  Applications. 

obtained  by  installing  several  pumps  in  series.  The  diagram 
should  be  studied  in  connection  with  Sees.  95  to  105.  The 
letters  S  and  D  refer  to  single  or  double-action.  The  type 
names  which  are  underscored,  indicate  the  type  ordinarily 
preferable  for  the  stated  conditions. 

107.  To  Compute  The  Horse  Power  Rating  Which  A  Motor 
Should  Have  to  Operate  A  Deep-Well  Pump  use  the  following 
formulas  which  were  derived  from  data  in  the  Goulds  Mfg. 
Co.  catalogue. 

When  the  pump  operates  single-acting  (Fig.  101)  or  two- 
stroke  (Figs.  103  and  104) : 

(48)  Pbhp  =  -?OQQ^  (horse  Power) 


98 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  3 


When  the  pump  operates  double-acting  (Fig.  102) : 


(49) 


Vam(L 


amhmT 


LfK) 


2000 


(horse  power) 


Wherein :  Pbhp  =  the  required  horse  power.  V om  —  the  quan- 
tity of  water  pumped  in  gallons  per  minute.  LhmT  =  the  total 
measured  head  against  which  the  pump  works  in  feet.  L,  = 
the  length  of  the  plunger  rod  in  feet.  K  =  a  constant  taken 
from  Table  108  by  which  the  weight  of  the  plunger  rods  and 
couplings  is  included  in  the  computation. 

NOTE. — The  quantity  K  is  ignored  in  For.  (48)  because  the  weight  of 
the  single-acting  rods  which  have  to  stand  tension  only  is  not  great  enough 
to  enter  into  the  calculation.  The  two-stroke  pump  plunger  rods  bal- 
ance as  explained  in  Sec.  90. 

EXAMPLE. — A  two-stroke  deep- well  pump  (Fig.  103)  is  required  to 
deliver  100  gal.  of  water  per  minute  against  a  total  measured  head  of 
200  ft.  What  should  be  the  horse-power  rating  of  a  motor  which  is  to 
drive  this  pump? 

SOLUTION.— By  For.  (48),  Pbhp  =  VgmLhmT/13W  =  100  X  200  -5- 
1300  =  15  h.p.  approximately. 

EXAMPLE. — A  double-acting  single-plunger  pump  is  required  to  draw 
125  gal.  per  min.  of  water  from  a  well  150  ft.  deep  and  deliver  it  into  a 
tank  100  ft.  above  ground.  The  bore  of  the  pump  cylinder  is  5.75  inches. 
What  should  be  the  horse-power  rating  of  a  motor  to  drive  this  pump? 

SOLUTION. — The  total  measured  head  =  160  +  100  =  250  ft.  By 
Table  108,  the  value  of  K  for  a  5.75  inch  pump  =  0.56.  By  For.  (49) 
Pbhp  =  Vgm(LhmT  +L/.K)  72000  =  125  X  [250  +  (150  X  0.56)]  -s-  2000  = 
20.9  h.p.,  or  21  h.p.  practically. 

108.  Table  Of  Head-Pressure  Equivalents  K  For.  (49)  Of 
Weight  Of  Deep-Well  Pump  Plunger  Rods. 


Dia.  pump 
cyl.  inches 

K  or  head  per 
ft.  of  rod 

Dia.  pump 
cyl.  inches 

K  or  head  per 
ft.  of  rod 

2.25 

0.96 

4.75 

0.59 

2.75 

0.69 

5.75 

0.56 

3.25 

0.72 

6.50 

0.46 

3.75 

0.73 

7.50 

0.36 

4.25 

0.60 

8.50 

0.40 

109.  Leather  Cup -Washers  For  Deep -Well  Pump -Plungers 
should  be  of  the  best  quality  of  oak  tanned  leather.  Soft 
spongy  leather  is  utterly  unsuited  for  this  service. 


SEC.  110] 


CRANK-ACTION  PUMPS 


99 


dv 


NOTE. — Leather  packing  should  be  thoroughly  greased  with  pure  tal- 
low. The  tallow  should  be  worked  into  the  leather  with  the  fingers  be- 
fore the  cup  is  put  into  place.  Satisfactory  lubrication  may  also  be 
secured  by  soaking  the  cups  in  neatsfoot,  sperm,  or  castor  oil  for  an  hour 
before  putting  them  into  place.  In  no  case  should  mineral  oil  be  used. 
Treatment  with  ordinary  machine  oil, 
which  contains  a  mineral  ingredient, 
tends  to  rot  the  leather  and  render  it 
pulpy. 

NOTE.— To  MAKE  A  SET  OF  CUP- 
WASHERS  FOR  A  PUMP  PLUNGER, 
proceed  as  shown  in  Fig.  119.  The 
cast-iron  mould,  M,  should  be  made 
with  d\  equal  to  the  diameter  of  the 
pump  cylinder  and  S  %2  in.  greater 
than  the  thickness  of  the  leather, 
(Table  110).  The  radius  of  the 
mould  at  R  should  be  about  one 
third  the  height  of  the  washer.  A 

disk  of  leather,  the  proper  diameter  and  thickness,  is  soaked  in  water  until 
soft.  Then  it  is  drawn  down  slowly  into  shape  by  means  of  the 
bolt.  The  protruding  edge  is  then  trimmed  off  flush  with  the  matrix. 
After  ten  hours  or  more,  the  leather  is  removed  and  well  greased  with 
tallow. 

110.  Table  Of  Dimensions  Of  Cup-Washers  For  Pump 
Plungers. 


FIG.  119.— Mold    For    Forming    Cup- 
Leathers. 


Diameter  of  pump 
cylinder  in  inches 

Thickness  of 
leather  in  inches 

D 

Depth  of  cup 
in  inches 

2 

Ke 

y* 

3 

Me 

H 

4 

H 

i 

5 

H 

IX 

6 

H 

IH 

QUESTIONS  ON  DIVISION  3 

1.  What  are  the  two  principal  classes  of  crank-action  pumps?     Define  each. 

2.  Why  may  steam  be  used  expansively  in  crank-and-fly-wheel  pumps  and  not  in 
direct-acting  pumps? 

3.  Give  values  for  the  steam  consumption  of  high-duty  crank-and-fly-wheel  pumps, 
run  condensing.     Non-condensing. 

4.  What  are  the  disadvantages  of  crank-and-fly-wheel  pumps,  as  compared  to  direct- 
acting  steam  pumps? 


100  STEAM  POWER  PLANT  AUXILIARIES  [Div.  3 

5.  What  two  kinds  of  deep  well  pumps  force  water  up  the  drop  pipes  in  a  fairly 
continuous  stream?     What  kind  does  not?     Can  this  last  kind  be  made   to  give  fairly 
continuous  discharge?     How? 

6.  Explain,  with  a  sketch,  the  operation  of  a  double-acting  single-plunger   deep-well 
pump.     Of  a  two-stroke  pump. 

7.  What  do  the  graphs  of  Figs.  112,  113  and  115  represent?     What  do   they  show 
about  the  action  of  various  kinds  of  pumps? 

8.  Under  what  condition  can  alcohol  and  gasoline  be  pumped  satisfactorily? 

9.  Give  several  advantages  of  electric  drive  for  a  remotely  located  pump. 

10.  What  type  of  pump  is  superseding  the  large  crank-and-fly- wheel  pump?     Why? 

11.  What  is  the  advantage  of  the  simplex  double-acting  pump  for    small  capacity 
requirements? 

12.  Name  two  widely-used  types  of  crank-action  power  pump  other  than  the  simplex 
double-acting  type.     Which  of  the  two  is  most  commonly  used? 

13.  Outline  a  method  of  arriving  at  a  proper  choice  of  power  pump. 

14.  What  is  a  cup  washer  for?     Explain  by  a  sketch  how  to  make  one.     How  should 
it  be  lubricated? 

PROBLEMS  ON  DIVISION  3 

1.  Compute  the  proper  horsepower  rating  for  a  motor  which  is  to  drive  a  single- 
acting  deep-well  pump  delivering  150  gal.  per  min.  against  a  total  measured  head  of 
225  ft. 

2.  Compute  the  proper  horsepower  rating  for  a  motor  which  is  to  drive  a  double- 
acting  deep-well  pump  having  a  displacement  of  0.9  gal.  per  rev.  at  30  r.p.m.     The 
well  rod  is  175  ft.  long  and  the  pump  discharges  50  ft.  above  the  base  of  the  generating 
gear.     Cylinder  diameter  is  2%  in. 


DIVISION  4 


CENTRIFUGAL  AND  ROTARY  PUMPS 

111.  The  Development  Of  The  Centrifugal  Pump  started 
with  its  invention  in  about  1680.  The  first  centrifugal  pump 
built  in  America  (Fig.  120)  was  called  the  Massachusetts 
pump.  This  was  a  crude  affair  of  low  efficiency.  Only 
during  the  last  20  years  has  much  improvement  been  made 
over  the  Massachusetts  pump.  This  seemingly  slow  develop- 
ment has  been  due  to  the  fact 
that  the  centrifugal  pump  is 
inherently  a  relatively -high- 
speed machine.  Formerly,  there 
was  no  motive  power  well  adapted 
to  drive  it.  The  introduction  of 
the  electric  motor  and  the  steam 
turbine,  which  are  inherently 
high-speed  machines,  led  to 
further  development.  Hence  the 
demand  for  centrifugal  pumps 
is  now  great  and  is  steadily 
increasing. 


FIG.    120.  —  The    Massachusetts 
Pump. 


NOTE. — A  large  portion  of  the  material  contained  in  this  Div.  is  based 
on  that  from  publications  of  The  Goulds  Manufacturing  Co.,  to  whom 
credit  is  hereby  given. 

112.  A  Centrifugal  Pump  is  a  pump  that,  as  will  be  ex- 
plained later,  depends  upon  centrifugal  force  or  the  variation 
of  pressure  due  to  rotation  for  its  action.     When  any  body 
is  constrained  to  move  in  a  curved  path,  there  is  a  force  which 
tends  to  impel  the  body  outward  from  the   center.     This 
force  is  called  centrifugal  force. 

113.  The  Theory  Of  The  Centrifugal  Pump  may  be  illus- 
trated (Figs.  121  and  122)  by  the  phenomenon  of  a  bucket  of 
water   which  is  whirled  around  the  head  in  a  circular  path. 
If  the  bucket  of  water  is  whirled  at  a  sufficiently-high  speed, 

101 


102 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


none  of  the  water  will  spill,  even  when  the  bucket  is  in  the 
position  shown  in  Fig.  121.  The  force  which  holds  the  water 
against  the  bottom  of  the  bucket  is  centrifugal  force.  Now  if 


Direct /'on  Of  Centrifugal  force  Is  Rcto/ict/fy 
..-•Out 'warot 'From  Center 


Circular  Path  Of  / 

Bucket-^  / 


FIG.  121.  —  Illustrating  Centrifugal  Force.       FIG.  122.—  Centrifugal  Force  Holds  The 

Water  Against  Bottom  Of  Bucket. 


Direction  Of  Centrifugal  Force  Is  Rao/- 
^  Jalty  Outwaro/  From  Center-, 


\   PapicJIy         \ 
'•Whir -/ina  \ 

ducket  N 


FIG.  123.— Illustrating  The  Principle  Of  FIG.  124.— Showing  That  Centrifugal 

The  Centrifugal  Pump.  Force  Causes  The  Water  To  Flow  Out- 

ward Through  The  Hole  In  The  Bucket. 

a  hole  is  cut  in  the  bottom  of  the  bucket,  the  water  will  be 
forced  out  through  the  hole  (Figs.  123  and  124)  and  will  be 
thrown  upward  into  the  air. 


SEC.  114]        CENTRIFUGAL  AND  ROTARY  PUMPS 


103 


Sucthn  P/'pe(Boys     \ 
Arm)..  \ 


EXPLANATION. — Suppose  that  the  boy's  arm  is  a  suction  pipe  and  that 
his  body  is  a  reservoir  containing  water  (Fig.   125).     The  centrifugal 
force  of  rotation  throws  the  water  from  the  bucket.     This  tends  to  pro- 
duce a  vacuum  within  the  bucket  and 
suction   pipe.     If    the    surface    of    the     ,-water  Thrown^   .  ,f     ,e     ^-^ 

.       f,  .      .  ,,        <FrvmBud(ef-/',<vit'r'fu9farr°rc*       ^ 

water  in  the  reservoir  is  open  to  the    \ £J  _      \ 

atmosphere,  the  water  will  be  forced  to 
rise  in  the  suction  pipe  by  the  atmos- 
pheric pressure  and  will  be  pushed  by 
the  centrifugal  force  out  through  the 
hole  in  the  bucket  so  long  as  the  end 
of  the  pipe  (Fig.  125)  is  submerged  in 
the  water  and  the  bucket  is  rotated  at 
a  sufficiently-high  speed.  Roughly, 
this  illustrates  the  theory  of  the  cen- 
trifugal pump. 


114.  The  Commercial  Cen- 
trifugal Pump  (Fig.  126)  is 
merely  a  modification  of  the  ap.- 
paratus  shown  in  Fig.  125.  The 


impeller,  rotating  within  the  cas-    FlG.  125!—  illustrating  The  Principle 


ing,  C,  corresponds  to  the  rotating          Of  The  Centrifugal  Pump. 
bucket.     Water   enters   the    impeller  through   an  inlet  hole 
around  its  center,  0.     The  rotation  of  the  impeller  imparts  cen- 
trifugal force  to  each  particle  of  water,  which   causes  the 


Suction— 
In/cf 


I-  Section  AA 


FIG.   126.  —  Single-Stage,  Single-Suction  Volute  Centrifugal  Pump. 

water  to  be  thrown  outward.  Thereby  pressure  is  created 
back  of  each  particle  of  water  and  the  water  is  discharged 
from  the  impeller  into  the  case,  C.  The  contour  of  the  im- 


104 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


peller  blades  is  so  designed  that  the  water  enters  the 
blades,  passes  through  them  and  is  discharged  with  a  minimum 
of  friction. 

EXPLANATION. — The  water  upon  entering  the  pump  at  0,  (Fig.  127) 
is  caught  between  the  vanes  of  the  impeller  which  are  rotating.  This 
rapid  rotation  of  the  water  sets  up  a  centrifugal  force,  F,  and  forces  the 
water  outward  against  the  pump  casing,  C,  just  as  the  boy  swinging  the 
bucket  over  his  head  (Fig.  121)  created  a  centrifugal  force  which  pressed 
the  water  against  the  bottom  of  the  bucket.  The  pressure  which  is 

^^^—^'Scrxyrge  Pipe  I  SqJn.  In  Cross  Section 

Water  Enters  Pump  Through  This 

\     ,--Sucthn  Pipe 
1+.. — Nozzle  y 

Direction  Of-. 
Retortion    • 


FIG.   127. — Centrifugal  Force  Created  By  Rotation  Of  Impeller  Vanes. 

thus  set  up  may  be  imagined  to  be  transmitted  by  the  water,  from 
particle  to  particle,  entirely  around  the  inner  periphery  of  the  casing  to 
the  discharge  nozzle.  The  water  is  thus  caused  to  rise  in  the  dis- 
charge pipe  P,  just  as  the  water  was  forced  out  through  the  hole  in  the 
bottom  of  the  rotating  bucket.  The  water  will  rise  in  the  pipe  until  the 
pressure  due  to  the  water  column  in  P  just  balances  the  centrifugal  force 
F.  Suppose  the  speed  of  rotation  of  the  impeller  7,  (Fig.  127)  is  such  that 
a  centrifugal  force  of  43.4  Ib.  per  sq.  in.  is  produced  on  the  casing.  Sup- 
pose the  nozzle  and  discharge  pipe,  P,  have  a  cross-sectional  area  of 
1  sq.  in.  Water  will  then  rise  in  the  discharge  pipe  until  the  weight  of 
the  water  column  is  43.4  Ib.  The  height  of  a  water  column  1  sq.  in.  in 
cross  section  having  a  weight  of  43.4  Ib.  is  (Sec.  5)  100  ft.  It  will  be 


SEC.  115]         CENTRIFUGAL  AND  ROTARY  PUMPS 


105 


shown  later  that  the  impeller  velocity  which  is  required  to  lift  water 
vertically  100  ft.  is  the  same  as  that  velocity  which  the  water  would  have 
after  freely  falling  through  a  distance  of  100  ft. 

NOTE. — THERE  ARE  OTHER  FACTORS  WHICH  HAVE  CONSIDERABLE 
EFFECT  upon  the  efficient  operation  of  centrifugal  pumps,  such  as  elimi- 
nation of  eddy  currents,  efficient  transformation  of  kinetic  energy  to 
pressure  without  shock,  etc.  These  are  principles  of  design  and  are 
not  within  the  scope  of  this  book. 

115.  A  Freely -Falling  Body  Will,  If  It  Falls  Through  A 
Certain  Height,  Have  A  Certain  Velocity,  or  speed,  at  the  end 
of  its  fall.  Suppose  there  is  a  body,  say 
a  bucket  of  water,  on  the  top  of  a  building 
(Fig.  128)  which  is  100  ft.  high.  If  the 
bucket  is  pushed  off  and  allowed  to  fall, 
it  will  fall  with  a  continuously  increasing 
speed  until  it  strikes  the  earth.-  If  it  is 
now  impelled  upward  with  an  initial 
velocity  equal  to  the  velocity  which  it 
had  when  it  struck  the  earth,  it  will 
rise  just  to  the  height  from  which  it  fell. 
The  velocity  which  a  body  will  acquire 
in  falling  through  a  given  distance,  or 
the  velocity  which  must  be  imparted  to  FIO  128_Vessel  Fall. 
a  body  to  cause  it  to  rise  to  a  given  ing  From  Top  of  A  ioo-Ft. 
height  may  be  computed  by  the  follow-  Buildine- 
ing,  which  is,  if  the  frictional  resistance  of  the  air  be  disre- 
garded, true  for  any  body  whatsoever: 

(50)  v  =  A/20L/  (ft.  per  sec.) 
or 

(51)  vm  =  481\/Lf  (ft.  permin.) 

Wherein:  v  =  velocity  in  feet  per  second.  vm  =  velocity  in 
feet  per  minute,  g  =  acceleration  due  to  gravity  =  32.2  ft. 
per  sec.  per  sec.  Lf  =  distance  in  feet,  through  which  body  falls, 
or  the  height  to  which  it  will  rise  if  impelled  upward  with  an 
initial  velocity  of  v  or  vm. 

EXAMPLE. — A  vessel  of  water  (Fig.  128)  is  dropped  from  a  point  100 
ft.  above  the  earth.  With  what  velocity  will  it  strike  the  ground? 
SOLUTION.— By  For.  (51),  the  velocity  =  vm  =  48lVLf  =  481\/TOO  = 
481  X  10  =  4,810  ft.  per  min. 


106 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


EXAMPLE. — What  is  the  initial  velocity  which  must  be  imparted  to 
the  vessel  of  water  to  cause  it  to  rise  100  ft.  in  a  vertical  direction? 
SOLUTION.— By  For.  (51),  the  velocity  =  vm  =  481\XZ/  =  48lVlOO  = 
481  X  10  =  4,180#.  per  min. 

116.  The  Theoretical  Speed  In  R.P.M.  At  Which  A  Cen- 
trifugal Pump  Impeller  Must  Run  To  Pump  Water  To  A  Cer- 
tain Height  may  be  determined  by  The  Law  Of  Freely  Falling 

Bodies.  As  was  shown  in  the 
preceding  Sec.,  the  water,  to  be 
thrown  to  a  certain  height,  must 
have  the  same  velocity  when  it 
leaves  the  impeller  as  it  would 
have  if  it  fell  from  the  same 
height.  This  may  be  stated: 
The  speed  in  feet  per  minute  of 
a  point  on  the  periphery  of  the 
impeller  should  be  equal  to  the 
velocity  which  the  water  would 
acquire  in  falling  from  the  sam# 
height  as  the  total  head  pumped 
against. 

NOTE. — THE  TOTAL  HEAD  PUMPED 
AGAINST  is  the  sum  of  all  friction, 
velocity,  and  static  heads,  which  occur 
between  the  suction-pipe  intake  and 
the  delivery-pipe  outlet.  See  Sec.  12. 
EXAMPLE. — At  what  speed,  in  r.p.m. 
must  a  12-in.  diameter  impeller  of  a 

FIG.  129.— A  12-in.  Diameter  Im-  centrifugal  pump  (Fig.  129)  be  driven 
peller  when  Driven  At  1685  R.P.M.  to  deliver  water  against  a  total  head 

^Theoretically,  Produce  A  12  Qf    m    f t  ?      ^idfon.— By   For     (51), 

velocity  =  vm  =  481\/L/  =481\/121  = 

481  X  11  =  5,291  ft.  per  min.,  which  is  the  required  peripheral  velocity 
of  the  impeller.  Circumference  of  impeller  =  -n-d  =  3.1416  X  1  =  3.14  ft., 
which  is  the  distance  a  point  on  the  periphery  of  the  impeller  will  travel 
during  1  revolution.  Now,  3.14  X  r.p.m.  =  peripheral  velocity  of  the 
impeUer  =  5,291.  Or,  r.p.m.  =  5,291  -T-  3.14  =  1,685  r.p.m. 

NOTE. — Due  to  certain  losses  which  cannot  be  eliminated,  the  actual 
speed  of  the  impeller  must  be  somewhat  greater  than  the  theoretical 
speed  to  produce  a  given  head. 

NOTE. — "HEAD"  MAY  BE  REDUCED  To  EQUIVALENT  POUNDS  PER 
SQUARE  INCH  UNIT  PRESSURE  as  explained  in  Sec.  4.  Also  see  the 
author's  PRACTICAL  HEAT  for  definition  and  explanation  of  unit  pressure. 


SEC.  117]         CENTRIFUGAL  AND  ROTARY  PUMPS  107 

117.  The  Quantity  Of  Water  Which  A  Pump  Will  Deliver 

when  being  driven  at  a  given  speed  will  depend  upon:  (1)  The 
size  of  the  discharge  outlet.  (2)  The  size  of  the  suction  inlet.  (3) 
The  size  of  the  casing.  (4)  The  width  of  the  impeller  vanes.  In 
good  design  the  allowable  velocity  of  the  water  at  the  discharge 
outlet  is  about  10  ft.  per  sec.  However,  this  velocity  may 
vary  from  5  to  15  ft.  per  sec. 

NOTE. — IT  Is  CUSTOMARY,  IN  ORDINARY  PARLANCE,  To  SPEAK  OF  A 
CENTRIFUGAL  PUMP  As  A  "4-w.  pump,"  a  "6-in.  pump,"  ETC.  This 
means  that  the  inside  diameter  of  the  discharge  nozzle,  N,  Fig.  126,  is 
4  in.  or  6  in.  However,  the  discharge-nozzle  diameter  is  not  to  be  taken 
as  accurately  denning  the  capacity  of  a  pump.  But  if  it  is  remembered 
that  the  nozzle-velocity  in  most  centrifugal  pumps  is  about  10  ft.  per 
sec.,  the  discharge-nozzle  diameter  does  provide  some  idea  as  to  the 
capacity  of  the  pump  in  gallons  per  minute.  An  approximate  rule  is: 
The  number  of  gallons  discharged  per  minute  is  approximately  equal  to  the 
square  of  the  discharge-nozzle  diameter,  in  inches,  multiplied  by  25. 

118.  The  Quantity  Of  Water  Delivered  By  A  Centrifugal 
Pump  Through  A  Frictionless  Pipe  Will  Vary  In  Direct  Pro- 
portion To  The  Speed  Of  The  Impeller,  If  The  Diameter  of  the 

impeller  remains  unchanged,  and  if  the  friction  of  the  water  in 
the  pump  is  neglected.  This  may  be  formulated  as  follows: 

(52)  Vam*  =   N*  j^7"""  (gal.  per  min.) 

Wherein :  Vgmz  =  quantity  of  water,  in  gallons  per  minute, 
delivered  by  the  pump  when  running  at  N2  r.p.m.  Vgm\  = 
quantity  of  water  delivered  by  the  pump  when  running  at 
Ni  r.p.m. 

EXAMPLE. — A  certain  centrifugal  pump  running  at  1,600  r.p.m.  de- 
livers 1,000  gal.  per  min.  through  a  frictionless  pipe  line.  How  many 
gallons  will  be  delivered  per  minute  by  the  same  pump  through  the  same 
pipe  if  the  speed  is  changed  to  1,200  r.p.m.  SOLUTION. — By  For.  (52), 
the  quantity  which  will  be  delivered  at  the  changed  speed  =  Vgmz  =  (Nz  X 
Vami)-/Ni  =  (1,200  X  1,000)  -5-  1,600  =  750  gal  per  min. 

NOTE. — SINCE  ALL  ACTUAL  PIPE  LINES  OFFER  FRICTIONAL  RESIST- 
ANCE To  WATER  FLOW  IN  THEM,  THE  ABOVE  FORMULA  CANNOT  BE 
USED  IN  PRACTICE.  The  actual  quantity  of  water  delivered  by  a  pump 
through  a  pipe  line  may  be  either  greater  or  less  than  the  value  obtained 
by  applying  the  above  formula.  The  only  practical  method  of  deter- 
mining the  delivery  of  an  actual  pump  at  different  speeds  is  by  test,  as 
explained  in  Sec.  138. 


108  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

119.  The  Pressure  Head  Which  Will  Be  Produced  By  A 
Centrifugal  Pump  Will  Vary  As  The  Square  Of  The  Speed  Of 
The  Impeller,  if  the  diameter  of  the  impeller  remains  constant 
and  there  is  no  water-friction  loss  within  the  pump.  This  may 
be  expressed  as  a  formula  by: 

(53)  LhT2  =  (^)  !,*„  (feet) 

Wherein:  LAr2  =  head,  in  feet,  produced  by  the  pump  when 
running  at  N2  r.p.m.  Lhri  =  head,  in  feet,  produced  by  the 
pump  when  running  at  Ni  r.p.m. 

EXAMPLE.  —  A  pump  which  has  no  water-friction  loss  is  running  at 
1,600  r.p.m.  produces  a  total  head  of  80  ft.  What  head  will  be  produced 
by  the  same  pump  if  the  speed  of  the  impeller  is  changed  to  1,000  r.p.m.? 
SOLUTION.  —  By  For.  (53),  the  head  produced  at  the  new  speed  =  LhTz  = 
(N2  -r  NJ*  X  LhTi  =  (1,200  -T-  1,600)2  X  80  =  %6  X  80  =  45  ft. 


120.  The  Power  Required  To  Drive  A  Centrifugal  Pump 
Will  Vary  As  The  Cube  Of  The  Speed  Of  The  Impeller,  if 
the  diameter  of  the  impeller  remains  unchanged,  and  if  no 
power  is  lost  through  pump  by  mechanical  and  water  friction. 
This  rule  may  be  written: 

(54)  Pbhp2  =  \~\  Pbhpl  (horse   power) 

Wherein  :  PbhP2  =  horse  power  required  to  drive  the  pump  at  a 
speed  of  N%  r.p.m.  PbhPi  =  horse  power  required  to  drive  the 
pump  at  a  speed  of  NI  r.p.m. 

EXAMPLE.  —  32  h.p.  are  required  to  pump  a  given  quantity  of  water 
against  a  certain  head  when  the  frictionless  pump  is  running  at  1,600 
r.p.m.  What  would  be  the  horse  power  required  to  drive  the  same  pump 
at  1,200  r.p.m.?  SOLUTION.  —  By  For.  (54),  the  power  required  at  the 
new  speed  =  Pbhpz  =  (N2  -5-  #i)3  X  Pbhpi  =  (1,200  -i-  1,600)  8  X  32  = 
27/64)  X  32  =  13.5  h.p. 

121.  The  Velocity  Of  A  Point  On  The  Periphery  Of  The 
Impeller  Is  Directly  Proportional  To  The  r.p.m.   Of    The 
ImpeUer,  or  expressed  as  a  formula: 

(55)  Vrn  =  N        d  =  0.261,8  Nd          (ft.  permin.) 


SEC.  122]         CENTRIFUGAL  AND  ROTARY  PUMPS  109 

Wherein :  vm  =  velocity,  in  feet  per  minute,  of  a  point  on  the 
periphery  of  the  impeller.  N  =  speed,  in  r.p.m.,  of  the 
impeller,  d  =  diameter  of  the  impeller  in  inches. 

NOTE. — By  transposing  For.  (55)  and  substituting  in  Fors.  (52),  (53), 
and  (54),  there  results: 
From  For.  (52) 

(56)  Vom*  =  dz  XVaml  (gal.  per  min.) 

«i 

From  For.  (53) 

(57)  LhT,  =    (~^j  2  LhTl  (feet) 
And  from  For.  (54) 

(58)  Pbhpz  =   (-A    Pbhpi  (horse  power) 

Wherein:  di  and  d2  =  the  old  and  new  diameters  of  the  impeller,  in 
inches,  respectively.  From  Fors.  (56),  (57),  and  (58),  it  is  evident,  that 
if  the  speed  in  r.p.m.  of  a  centrifugal-pump  impeller  remains  constant, 
and  if  there  is  no  friction,  the  following  will  be  true:  (A)  From  For.  (56), 
the  quantity  of  water  delivered  will  vary  as  the  diameter  of  the  impeller.  (B) 
From  For.  (57),  the  head  produced  will  vary  as  the  square  of  the  impeller 
diameter.  (C)  From  For.  (58),  the  power  required  for  driving  will  vary  as 
the  cube  of  the  impeller  diameter. 

122.  Centrifugal  Pumps  May  Be  Classified  According  To 
Several  Different  Features,  the  most  important  of  which  are: 
(1)  Volute  or  turbine.     (2)  The  number  of  stages.     (3)  Single 
suction    or    double   suction.     (4)    Open    impeller   or   enclosed 
impeller.     (5)  Horizontal  or  vertical.     Each  of  these  different 
features  will  be  discussed  in  succeeding  Sees. 

123.  The    Two    General    Classifications    Of    Centrifugal 
Pumps  Are:  (1)  Turbine  Pumps.     (2)  Volute  Pumps.    The 
turbine  pump   (Fig.    130)  is  one  wherein  the  impeller  is  sur- 
rounded by  a  diffusor  containing  diffusion  vanes  which  direct 
the  water  flow  from  the  impeller.     The  relative  position  of 
the  diffusor,  D,  and  the  diffusion  vanes,  V,  (also  called  guide 
vanes)  is  shown  in  Fig.  131.      These  vanes  are  so  shaped  that 
gradually   enlarging   passages   are   provided   for   the   water. 
In  flowing  through  these   guide-vane  passages,  the  velocity 
which   is   imparted   to   the  water  by   the    centrifugal  force 
(Sec.  114)  is  converted  into  pressure.     The  casing  which  sur- 
rounds the  diffusion  ring,  may  be  circular  and  concentric 


110 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 


SEC.  124]         CENTRIFUGAL  AND  ROTARY  PUMPS 


111 


(Fig.  131)  with  the  impeller,  or  is  sometimes  of  a  spiral  form. 
The  volute  pump  (Fig.  126) 
is  one  which  has  no  guide 
vanes,  but  instead,  has  a 
spiral-shaped  casing.  This 
spiral  casing  F,(Fig.  126) 
is  also  called  the  volute. 
In  the  volute  pump,  this 
spiral  casing  replaces  the 
guide  vanes  of  the  turbine 
pump.  The  volute,  or 
spiral  casing,  is  so  de- 
signed that  it  so  guides  the 
water  from  the  impeller  to 
the  discharge  pipe  that  the 
velocity  is  gradually  con- 
verted into  pressure.  Vol- 
ute pumps  ordinarily  have 
but  a  single  impeller. 
Where  a  closed-type  im- 
peller is  used,  a  double- 
inlet  is  employed,  thereby 
eliminating  end  thrust. 

124.  The  Applications  Of 
The  Volute  Pumps  And  Of 
The  Turbine  Pumps  over- 
lap. In  general,  however, 
for  low  heads,  (under  about 
70  or  80  ft.)  the  volute 
pump  should  be  chosen. 
For  higher  heads  the  tur- 
bine (multi-stage,  Sec. 
126)  pump  will  give  better 
service.  The  volute  pump 
may  be  considered  superior 
to  the  turbine  pump  from 
the  standpoint  of  size,  sim- 
plicity, and  cheapness. 


112 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


NOTE. — THERE  Is  MUCH  CONTROVERSY  CONCERNING  THE  COMPARA- 
TIVE EFFICIENCY  OF  THE  Two  TYPES  OF  PUMPS.  More  rapid  progress 
has  probably  been  made  in  the  design  of  the  turbine  pump  than  in  that 
of  the  volute  pump.  This  is  attributed  to  the  fact  that  the  guide-vane 
design  of  the  turbine  pump  is  more  amenable  to  mathematical  analysis 
than  is  the  spiral  casing  of  the  volute  pump.  It  has  been  predicted, 
that  volute  passages  will  eventually  be  designed  whereby  it  will  be  possi- 
ble to  effectively  pump  against  the  same  heads  with  the  volute  pump  as 
with  the  turbine  pump.  Since  the  volute  pump  is  the  cheaper  and 
simpler  it  may  therefore  find  a  wider  application  in  the  future. 

125.  Water  May  Be  Raised  As  High  As  Desired  by  arrang- 
ing a  sufficient  number  of  independent  pumps  (Fig.  132)  so 
that  the  discharge  of  one  of  the  pumps  is  piped  to  the  suction 

of  the  next.  It  is  desired  to 
pump  the  water  (Fig.  132)  to 
a  total  height  of  200  ft. 
Pump  A  takes  water  from 
reservoir  D  and  delivers  it 
to  reservoir  E.  Pump  B 
takes  water  from  reservoir 
E  and  delivers  it  to  reservoir 
F.  This  is,  however,  an  un- 
economical method  of  pump- 
ing water  against  a  high 
head.  The  usual  method 
which  is  used  in  practice  is 
described  in  the  following 
Sec. 

126.  The  Multi-Stage 
Centrifugal  Pump  (Figs.  130 
and  133)  is  really  two  or 
more  distinct  pumps  con- 
nected in  series.  Such  a 
pump  has  two  or  more  im- 
pellers through  which  the  water  passes  successively.  The 
impellers  are  mounted  on  the  same  shaft  and  contained  within 
the  same  casing.  That  is,  the  water  is  discharged  from  the 
first-stage  impeller,  7,  (Fig.  133)  through  the  return  chamber, 
Ri  to  the  suction  side  of  the  second-stage  impeller,  77,  etc., 
throughout  each  stage  of  the  pump.  Multi-stage  pumps  are 


FIG.  132. — Showing  How  Water  May  Be 
Pumped  To  A  Great  Height  By  Separate 
Steps  Or  Stages. 


SEC.  125]         CENTRIFUGAL  AND  ROTARY  PUMPS 


I 

113 


114  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

used   to  pump  against  high  heads.     They  may  be  either  of 
the  volute  or  of  the  turbine  type. 

EXPLANATION. — The  two-stage  pump,  (Fig.  130)  may  be  considered 
merely  as  a  more  compact  arrangement  of  the  two  pumps  in  Fig.  132. 
Suppose  the  water  is  taken  into  the  first-stage  suction,  Si  (Fig.  130)  and 
is  discharged  to  the  second-stage  suction,  S2,  through  the  return  chamber, 
R,  at  a  pressure  equivalent  to  a  100-ft.  head.  The  water  is  then  received 
by  the  second-stage  impeller  under  a  100-ft.  head.  In  passing  through 
the  second-stage  impeller,  the  water  is  given  an  additional  100-ft.  pres- 
sure head.  Thus  as  the  water  passes  from  Si  to  N  (Fig.  130)  the  same 
result  is  obtained  as  by  the  two  pumps  in  Fig.  132.  Multi-stage  pumps 
are  usually  designed  to  produce  from  about  a  100-  to  a  150-ft.  head  per 
stage.  The  superiority  due  to  compactness,  simplicity,  and  economy  of 
the  multi-stage  pump  of  Fig.  130  over  the  two-pump  arrangement  of 
Fig.  132  is  obvious. 

127.  "Single    Suction"    And    "Double   Suction"   are  also 
classifications  of  centrifugal  pumps.     A  single-suction    (also 
called  side-suction)  pump  (Fig.  133  is  one  in  which  the  water 
enters    the  impeller  from  one  side  only.     A  double-suction 
pump  (Fig.  130)  is  one  in  which  the  water  enters  the  impeller 
from  both  sides.     A  double-suction  pump  will,  with  same  im- 
peller diameter,  have  a  larger  discharge  than  a  single-suction 
pump.     The  double-suction  pump  may  have  two  separate 
suction  pipes,  or  the  water  may  be  divided  after  it  enters  the 
casing.     A  single-suction  pump  which  takes  water,  either  by 
suction  or  under  a  positive  head,  will  have  a  side-thrust.     Side- 
thrust  is  caused  by  the  pressure  on  one  side  of  the  impeller 
being  greater  than  the  pressure  on  the  other  side.     This  side- 
thrust  is  transmitted  to  the  shaft,  and  will,  unless  some  method 
of  balancing  is  provided,  cause  excessive  friction  and  wear  in 
the  thrust  bearing. 

128.  The  Forces  Which  Tend  To  Unbalance  The  Impeller 
may  be  understood  from  a  consideration  of  Fig.  134.     The 
water,  which  enters  the  impeller  eye  at  A,  has  its  direction  of 
flow  parallel  to  the  axis  of  the  shaft.     When  the  water  im- 
pinges on  the  impeller  at  B,  its  direction  of  flow  is  changed, 
as  shown  by  the  arrows.     This  change  of  direction  results 
in  the  exertion  of  a  force  against  the  impeller  which  tends  to 
move  it  to  the  right.     Since  the  pressure  in  pounds  per  square 
inch  in  r  is  almost  equal  to  the  pressure  in  pounds  per  square 


SEC.  129]         CENTRIFUGAL  AND  ROTARY  PUMPS 


115 


Cctslngr- 


inch  at  the  periphery  of  the  impeller,  the  water  in  r  will  exert 
a  force  on  the  impeller,  the  direction  of  which  will  be  to  the 
left.  Due  to  the  same  cause,  a  pressure  will  exist  in  t,  which 
will  exert  a  force  to  the  right  on  the  impeller.  However,  the 
leakage  of  water  through  s  will  result  in  the  pressure  in  pounds 
per  square  inch  in  t  being  somewhat  less  than  that  in  r.  Also, 
the  area  of  the  impeller  web  over  which  the  force  in  r  acts 
is  greater  than  that  over  which 
the  force  in  t  acts.  Therefore, 
since  the  pressure  in  pounds  per 
square  inch  in  t  is  less  than  that 
in  r,  and  since  the  area  of  r  is 
greater  than  that  of  t,  the 
combined  -  transmitted  -  pressure 
force  will  act  to  the  left  on  the 
impeller.  As  all  of  these  forces 
may  vary  from  one  instant  to 
the  next,  the  direction  of  the 
resultant  may  shift  from  right 
to  left.  It  cannot,  therefore, 
be  predetermined  just  how  great 
or  in  which  direction  the  result- 
ing force  will  be.  To  minimize 
the  total  resultant  unbalance 
the  devices  which  will  be  de- 
scribed are  employed. 

129.  There  Are  Various  Methods  Of  Balancing  Single- 
Suction  Impellers  Against  End -Thrust,  the  most  common  of 
which  are:  (1)  The  Jaeger  method.     (2)  By  means  of  an  auto- 
matic hydraulic  balancing  piston.     Each  will  be  described: 

130.  The    Jaeger    System    Of    Balancing    Single -Suction 
Impellers  (Fig.  135)  automatically  minimizes  the  longitudinal 
unbalance   but   it   requires,   in   addition,   mechanical   thrust 
bearings.     The  impeller  is  equipped  in  front  and  rear,  with 
wearing  rings  (R,  Fig.  135).     The  diameter  of  the  front  and 
back  rings  is  the  same,  so  that  the  area  of  the  surface  a  is 
equal  to  that  of  surface  b.     Since  leakage  through  the  rings 
will  be  practically  the  same  in  both  the  front  and  the  back 
sides,  the  pressure  on  a  will  be  equal  and  opposite  to  that  on 


FIG.  134. — Unbalanced  Impeller. 


116 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 


b.  The  leakage  water  which  flows  across  the  front  sealing 
surface  enters  the  suction  opening  of  the  impeller.  To  prevent 
the  leakage  water  which  flows  across  the  back  sealing  surface 
from  collecting  in  the  annular  ring  and  building  up  pressure, 
the  holes,  H  (Figs.  131  and  135),  permit 
this  leakage  water  to  pass  into  the  im- 
peller. Leakage  through  the  wearing 
rings  may  be  minimized  by  forming  a  laby- 
rinth pathway  (Fig.  136)  for  the  water. 
The  mechanical  thrust  bearings  which  are 
necessary  to  resist  the  force  due  to  change 
of  direction  (Sec.  128)  are  usually  of  the 
ball  type,  or  (Fig.  130)  of  the  multi-collar 
type. 

131.  The  Automatic  Hydraulic  Balanc- 
ing Piston  (Fig.  133)  whereby  all  of  the 
impellers  (multistage  pump)  are  balanced 
by  a  single  balancing  piston  is  shown  in 
Fig.  137.  This  balancing  chamber  is  at 
the  right-hand  end  of  the  last  stage.  The 
last-stage  impeller  is  provided  with  wearing 
rings.  Water  leaks  through  between  the  surfaces  of  these 
wearing  rings  to  the  balancing  chamber.  If  the  shaft,  and 
the  movable  part,  M,  (Fig.  137)  moves  to  the  right,  the  pas- 
sageway between  the  wearing  rings  is  increased.  This  permits 
the  water  to  pass  more  freely  through  R  into  the  balancing 


FIG.  135.  —  Showing 
Jaeger  Method  Of  Im- 
peller Balancing. 


f-Flat  type  H-Offset   Type  HT-Labyrinth.  Type 

FIG.   136. — Various   Types   Of   Wearing   Rings.      (7  =  Impeller,    W  =  Wearing    Ring, 

C  =  Casing.) 

chamber  C.  This  same  movement  to  the  right  tends  to  close 
the  escape-passageway,  E,  which  prevents  the  water  from 
escaping  through  the  pipe,  P.  Thus,  the  pressure  in  the 
balancing  chamber  builds  up  and  acts  against  the  balancing 


SEC.  132]         CENTRIFUGAL  AND  ROTARY  PUMPS 


117 


disk,  (or  piston)  D,  which  is  fixed  to  the  shaft.  This  moves 
the  shaft  to  the  left  until  R  is  closed  and'  E  is  open  and 
equilibrium  is  established. 

NOTE. — BALANCING  OF  DOUBLE-SUCTION  PUMPS  is  taken  care  of, 
theoretically,  in  the  design  of  the  pump.  The  liquid  is  supposed  to 
enter  in  equal  volumes  from  both  sides.  Since  the  inlet  openings  are 
also  supposed  to  be  equal,  the  vacuum  or  pressure  on  one  side  of  the 
impeller  is  always  equal  and  opposite  to  that  on  the  other  side.  There- 
fore, no  end-thrust  is  exerted.  The  impeller  is  also  equipped  with  front 
and  back  wearing  rings  of  equal  diameter  (Sec.  130)  so  that  there  is  no 
end-thrust  on  the  impeller  on  the  outside  of  the  wearing  rings.  Actually, 


L&Sf- 
Sfage 


Pipe  To  F/rst~Sf age,- Suction  Chctmbw 
•—Wearing  Rings 


FIG.  137. — Piston,   Or  Automatic,   Balancing  System   For  Centrifugal  Pumps. 
Laval  Steam  Turbine  Co.) 


(De 


however,  the  inlet-openings  are  never  exactly  equal.  The  wearing  rings 
are  likely  to  wear  unevenly.  One  or  both  of  these  causes  will  set  up  an 
unbalanced  end-thrust  on  the  inside  of  the  wearing  rings,  making  it 
necessary  to  equip  a  pump  of  this  type  with  a  mechanical  thrust  bearing. 
NOTE. — DUE  To  THE  SMALL  BEARING  SURFACE  OF  THE  OPEN-TYPE 
IMPELLER  (Fig.  127)  very  little  end-thrust  is  developed.  Hence,  mechan- 
ical thrust  bearings  will  ordinarily  assume  the  end-thrust  which  is  devel- 
oped in  a  pump  of  this  type. 

132.  The  Open  Impeller  is  shown  in  Figs.  138  and  138A. 
Pumps  equipped  with  an  impeller  of  this  type  are  sometimes 
called  fan  pumps.  The  action  is  similar  to  that  of  a  paddle- 
wheel  revolving  in  a  circular  casing.  All  of  the  early  centri- 
fugal pumps  were  of  this  type.  It  has  poor  water-guidance 
and  flow-lines.  This  results  in  excessive  wasteful  churning 
and  eddying  of  the  water.  Also,  a  great  amount  of  water 


118 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


escapes  between  the  blades  of  the  impeller  and  the  casing  walls. 
This  is  similar  to  the  slip  (Sec.  22)  in  reciprocating  pumps. 


Blade >• 


Boss-.. 


V    i    «     w  H-Sect  JonA-A 

FIG.   138. — Open  Type  Of  Impeller.     (Pumps  for  use  as  power-plant  auxiliaries  are 
seldom  equipped  with  open  impellers.) 

Due  to  the  above  mentioned  causes,  the  efficiency  of  the  pump 
which  is  equipped  with  an  open  impeller  is  comparatively  low. 
It  is  relatively  cheap  in  price.  For  certain  classes  of  work 


Fia.  138 A.— Perspective  View  Of  An  Open-Type  Centrifugal-Pump  Impeller. 

such  as  pumping  mash  and  thick  liquids,  it  is  the  only  type  of 
centrifugal  pump  that  will  give  satisfaction.  Its  use  is  not 
to  be  recommended  as  a  power-plant  auxiliary. 


SEC.  133]         CENTRIFUGAL  AND  ROTARY  PUMPS 


119 


133.  The  Enclosed  Impeller  (Fig.  139)  is  a  development  of 
the  open  impeller.  If  a  disk  or  plate  were  secured  to  each 
side  of  an  open  impeller,  a  closed  impeller  would  result.  The 
enclosing  walls  or  covers  are,  in  practice,  cast  solid  with  the 
impeller  vanes.  These  enclosing  walls  prevent  the  water 
from  escaping  past  the  impeller  blades.  Also  a  relatively 
close-running  joint  can  be  made  between  the  impeller  and  the 
casing.  This  reduces  the  slip  to  less  than  that  which  occurs 
with  the  open  impeller.  The  efficiency  of  the  pump  is  mate- 


Vcvne 


I-  Side  View  I-  Sectioned  Elevation 

FIG.   139. — Closed-Type  Impeller. 

rially  increased  by  these  two  devices.  The  running  joint  (Fig. 
135)  is  usually  known  as  the  sealing  surface.  The  running 
joint  is  formed  by  the  wearing  rings. 

NOTE. — THE  IDEAL  CONDITION  WOULD  BE  To  HAVE  A  TIGHT  FIT 
BETWEEN  THE  SEALING  SURFACES.  This  is,  however,  impossible  of 
attainment.  In  practise,  a  diametral  clearance  of  from  0.012  to  0.018 
in.,  is  allowed  between  the  wearing  rings.  Small  particles  of  grit  in  the 
water  will  cause  the  rings  to  wear,  thus  enlarging  the  clearance  and  in- 
creasing the  leakage.  The  increased  leakage  will  lower  the  efficiency. 
This  necessitates  renewing  of  the  wearing  rings. 

134.  The  Maximum  Heads  Against  Which  Impellers  Of 
The  Different  Types  Are  Designed  To  Operate  are  approxi- 
mately as  follows:  (1)  Single-suction,  open  impeller,  100  ft. 
(2)  Double-suction,  open  impeller,  100  ft.  (3)  Single-suction, 


120 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


Drir'ny 
Motor- 


enclosed  impeller,  100  ft.     (4)  Double-suction,  enclosed  impeller, 

150  ft. 

NOTE. — THE  SINGLE  IM- 
PELLER PUMP  (R.  A.  Fiske) 
may  be  efficiently  used  for 
heads  up  to  and  including  150 
ft.  or  higher,  with  efficiencies  of 
from  50  to  80  per  cent.  For 
pressures  above  50  Ib.  per  sq. 
in.,  two  or  more  runners  or 
stages  may  be  used,  each  stage 


adding  approximately  50  Ib.  per 
sq.  in.  to  the  total  pressure 
available  from  the  pump. 

135.  A  Vertical  -Shaft 
Centrifugal  Pump  (Fig. 
140)  may  be  used  where 
conditions  are  such  that  a 
horizontal-shaft  pump  can- 


j£     not  be  placed  within  suc- 


tion distance  of  the 
supply-  water  level.  This 
condition  is  frequently  en- 
countered in  deep  wells, 
sewage  service,  sumps, 
and  along  rivers  where  the 
difference  in  water  level 
between  high  and  low 
water  will  amount  to  20 
or  30  ft.  A  vertical  cen- 
trifugal pump  may  be 
operated  completely  sub- 
merged in  the  water  (Sec. 
136).  It  is,  however,  ad- 
visable, where  conditions 
permit,  to  locate  the  pump 
in  a  dry-pit.  This  makes 


FI0.    HO.-A 


O, 


than  when   submerged. 
Consequently  the  pump  will  be  given  better  attention.     For 


SEC.  136]         CENTRIFUGAL  AND  ROTARY  PUMPS 


121 


reasons,  which  will  be  stated  in  the  following  Sees.,  a  vertical- 
shaft  centrifugal  pump  should  not  be  selected  where  it  is 
feasible  to  use  one  of  the  horizontal-shaft  type. 

136.  The  Bearings  In  A  Vertical  Pump  are  very  likely  to  be 
a  source  of  constant  trouble.     These  bearings  may  be  divided 


Vertical  Line- 
-Shaft 


Drain  Plug- 


,     Oil     . 
-Reservoir 

^Adjustment  Bolt 


FIG.  141. — Sectional  View  Of  Hanger-Type 
Thrust  Bearing  For  Vertical  Centrifugal 
Pumps.  (Worthington  Pump  And  Machinery 
Corp.) 


FIG.  142.  —  Showing  Rotating 
Parts  And  Thrust  Bearing  Of  A 
Vertical  Centrifugal  Pump.  (The 
Goulds  Mfg.  Co.) 


into  two  classes:  (1)  The  pump  bearings  proper.  (2)  The  line- 
shaft  bearings.  The  pump  bearings,  if  the  pump  is  submerged, 
usually  depend  upon  the  water  for  lubrication.  This  results 
in  extremely  rapid  wear.  The  line-shaft  bearings  consist  of 
the  thrust  bearings  (Figs.  140,  141  and  142),  and,  if  the  line 
shaft  is  long,  the  guide  bearings  (Fig.  143).  The  thrust  bear- 
ing  must  carry  the  weight  of  the  rotating  parts  and,  in  some 
instances,  the  weight  of  the  pump.  It  has  been  found  difficult 


122 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


Vertical 
Shaft-' 


OH  Forced  Up 
Through  Pi'pp 
By  Centrifugal 

" 


to  design  a  thrust  bearing  which  will  operate  satisfactorily  at 
centrifugal-pump  speeds.  The  multi-collar  (Fig.  141), roller, 
and  self -aligning  ball  (Fig.  142)  types  of  bearings  are  used.  In 
any  event,  the  bearings  of  vertical  pumps  require  considerable 
attention. 

137.  When  The  Line  Shaft  Of 
A  Vertical  Pump  Is  Long,  It  Is 
Difficult  To  Keep  The  Motor, 
Line  Shaft  And  Pump  In  Align- 
ment. When  the  line-shaft 
length  exceeds  about  30  to  40 
ft.  a  certain  flexibility  and  the 
inevitable  misalignment  in  an 
installation  of  this  sort  must  be 
provided  for.  This  necessitates 
the  installation  of  several  thrust 
bearings  between  the  motor  and 
the  pump  with  a  flexible  coup- 
ling immediately  above  each 
thrust  bearing.  A  guide  bear- 
ing should  be  placed  on  each  side 
of  and  close  to  each  flexible  coup- 
ling. The  maximum  distance 
between  guide  bearings  should  not  exceed  about  6  ft. 

138.  The  Performance  Characteristics  Of  A  Centrifugal 
Pump  For  Various  Conditions  Of  Operation  should  be  known 
before  it  is  placed  in  any  given  installation.  The  factors  which 
determine  the  performance  characteristics  of  a  centrifugal 
pump  are  principally :  (I)  The  quantity  of  water  delivered.  (2) 
The  efficiency.  (3)  The  horse  power  input  at  each  of  several 
different  heads.  These  data  are  usually  supplied  by  the  manu- 
facturer, but  if  they  are  not,  they  may  be  secured  by  test.  The 
two  principal  reasons  for  testing  a  pump  are  :  (I)  To  determine 
its  characteristics  (Sec.  139).  (2)  To  determine  whether  or 
not  the  manufacturer's  guarantees  have  been  fulfilled. 

NOTE. — A  CENTRIFUGAL  PUMP  MAY  BE  TESTED  As  FOLLOWS:  The 
pump  which  is  to  be  tested  is  directly  connected  to  a  direct-current, 
variable-speed  electric  motor,  M  (Fig.  144)  of  known  efficiency.  A  volt- 
meter, V,  and  an  ammeter,  A,  are  connected  in  the  motor  circuit.  A 


I 

-Copper 
fates  With  Shaft 

FIG.  143. — Guide  Bearing  For  Verti- 
cal Shaft  Centrifugal  Pumps.  (Worth- 
ington  Pump  And  Machinery  Corp.) 


SEC.  138]        CENTRIFUGAL  AND  ROTARY  PUMPS 


123 


pressure  gage,  P,  is  connected  into  the  discharge  pipe.  A  vacuum  gage, 
S,  is  connected  into  the  suction  pipe.  The  quantity  of  water  discharged 
may  be  measured  either  by  means  of  a  calibrated  nozzle  placed  on  the 
end  of  the  discharge  pipe,  or  by  a  water  meter,  W.  Or,  if  the  pump  is  of 
small  capacity,  the  water  may  be  discharged  into  a  suitable  container,  R, 
and  measured  directly.  The  gage  readings  of  S  and  P  should  be  com- 
bined and  converted  to  head  in  feet  (Sees.  5  and  38),  which  will  be  the 
total  head  pumped  against  if  the  discharge-  and  suction-pipes  are  of  the 
same  diameter. 


/Suction  Gorge 


H^ 


_  _  — Ammeter 


Source  of  D.  C.  Supply 


Voltmeter 

{Discharge  Nozzle 
^^  Discharge  ReservolK^ 
J 


FIG.  144. — An  Arrangement  Which  May  Be  Used  In  Testing  A  Centrifugal  Pump. 


The  pump  is  primed  and  started.  The  speed  must  be  maintained 
constant  throughout  the  test.  Simultaneous  readings  of  S,  P,  A,  V',  and 
W  are  taken.  S  and  P  are  converted  into  total  head  in  feet  (Sees.  5  and 
38).  Then  thp«e  formulas  may  be  applied: 


(59) 
and 
(60) 


I  X  V  X  E, 
746 

LhT  X   Vgm 

39.6  X  Pbhp 


(horse  power) 
(per  cent.) 


Wherein:  Pbhp  =  input  to  pump  in  horse  power  (also  called  brake  horse 
power  of  pump).  /  =  motor-current,  in  amperes,  as  read  from  the 
ammeter.  V  =  motor-e.m.f.,  in  volts,  as  read  from  voltmeter.  Vgm  = 
quantity  of  water  pumped,  in  gallons  per  minute,  as  determined  from 
water  meter.  LhT  =  total  head  pumped  against,  in  feet,  as  obtained 
from  S  and  P.  ETO  =  efficiency  of  motor,  at  the  given  load,  expressed 
as  a  decimal,  as  obtained  from  the  motor  efficiency  graph.  Ep  =  effi- 
ciency of  the  pump,  expressed  in  per  cent. 


124  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

By  applying  the  formulas  for  a  certain  discharge  in  gallons  per  minute, 
the  head,  the  brake  horse  power,  and  the  efficiency  of  the  pump,  when 
running  at  the  given  speed,  are  determined.  The  discharge  is  now  varied 
by  either  opening  or  closing  the  gate-valve,  G,  and  another  set  of  readings 
is  taken  and  the  corresponding  computations  are  made  as  described 
above.  By  opening  or  closing  the  gate-valve,  the  conditions  should  be 
varied  from  no  discharge  when  G  is  closed,  to  practically  no  head  when  G 
is  wide  open.  Several  sets  of  readings  should  be  taken,  at  fairly  regular 
intervals  of  discharge  in  gallons  per  minute,  over  that  discharge  range 
which  will  be  provided  from  gate-valve  entirely  closed  to  wide  open.  The 
test  data  should  be  plotted  into  a  characteristic  chart  as  will  be  described. 

EXAMPLE. — A  centrifugal  pump  (Fig.  144),  which  is  undergoing  test, 
is  driven  by  a  direct-connected,  direct-current  motor,  at  a  constant  speed 
of  1,700  r.p.m.  A  certain  set  of  readings  are  as  follows:  Vgm  =  400  gal. 
per  min.;  S  =  8.9  in.  of  mercury;  P  =  26  Ib.  per  sq.  in.;  A  =  36  amp.; 
V  =  218  volts.  What  is  the  horse-power  input  to  the  pump,  the  head 
produced,  and  the  efficiency  of  the  pump  in  per  cent.,  at  this  rate  of 
discharge? 

SOLUTION. — By  note  subjoined  to  Sec.  38,  the  suction  pressure  developed 
=  (8.9  X  0.4914)  =  4.37  Ib.  per  sq.  in.  Since  the  water  level  is  below 
the  pump-center,  S  and  P  are  combined  by  addition,  or  (4.37  +  26)  = 
30.37  Ib.  per  sq.  in.  By  For.  (1),  the  total  head  produced,  LhT  =  (2.31  X 
30.37)  =  70  ft.  From  the  motor-characteristic  chart,  it  is  found  that 
the  efficiency  of  the  motor  at  this  load  ETO  =  89  per  cent.  By  For.  (59), 
the  horse-power  input  to  the  pump,  Pbhp  =  /XFXEOT-7-  746  =  36  X 
218  X  0.89  -3-  746  =  9.37  h.p.  By  For.  (60),  the  efficiency  of  the  pump, 
Ep  =  (LhT  X  Vom)  +  (39.6  X  Pbhp)  =  (70  X  400)  +  (39.6  X  9.37)  =66.2 
per  cent. 

NOTE. — A  CENTRIFUGAL  PUMP  SHOULD  BE  TESTED  UNDER  THE  CON- 
DITIQNS  To  WHICH  IT  WILL  BE  SUBJECTED  WHEN  INSTALLED.  Thus 
a  boiler-feed  pump  should  be  tested  with  water  at  the  temperature  of 
that  which  it  will  ultimately  handle. 


139.  A  Chart  Of  The  Characteristic  Graphs  Of  A  Centrifugal 
Pump  may  be  plotted  thus :  First  compute  from  the  test  data 
the  head  in  feet,  the  brake  horse  power,  and  the  efficiency,  for 
each  of  the  different  rates  of  discharge.  Then  (Fig.  145)  lay 
off,  on  the  horizontal  axis  (on  a  sheet  of  cross-section  paper),  of 
the  graph,  to  a  convenient  scale,  the  range  of  discharge  values 
in  gallons  per  minute.  Next  lay  off,  on  the  vertical  axis,  the 
range  of  values  corresponding  to  the  head  in  feet,  the  brake 
horse  power,  and  the  efficiency.  Now  plot  the  values:  Lay 
off  to  scale,  on  the  horizontal  axis,  distances  equivalent  in 
value  to  the  different  discharges  in  gallons  per  minute  as 


SEC.  140]         CENTRIFUGAL  AND  ROTARY  PUMPS 


125 


taken  from  the  test  data.  For  each  point  thus  obtained, 
locate  new  points  in  the  body  of  the  chart  by  laying  off  verti- 
cally, to  scale,  distances  which  are  equivalent  to  the  heads  in 
feet  for  the  discharge  at  each  head.  A  smooth  curve  drawn 
through  the  points  obtained  as  described,  results  in  the  head 
graph  (Fig.  145).  The  brake  horse  power  and  the  efficiency 
graphs  are  plotted  in  a  similar  manner.  These  three  graphs 
are  known  as  the  characteristics  of  the  pump. 


jgOO  2000 

Discharge    fn    Gall  6"n  s     Par     Minute 

Fia.  145. — Typical  Characteristics  Of  A  Centrifugal  Pump  At  Constant  Speed. 

140.  A  Number  Of  Important  Facts  May  Be  Determined 
From  The  Characteristic  Graphs  (Fig.  145)  of  a  pump  which 
is  operated  at  a  given  speed  which  are  not  apparent  from  the 
test  data,  such  as:  (1)  The  rate  of  discharge  in  gallons  per 
minute  when  pumping  against  any  head.  (2)  The  efficiency 
of  the  pump  at  any  discharge  rate.  (3)  The  horse  power  required 
to  drive  the  pump  when  pumping  water  against  any  head.  If 
(with  a  certain  pump  speed  in  r.p.m.)  any  one  of  the  four  items: 
the  head  pumped  against,  the  efficiency,  the  brake  horse 
power,  or  the  discharge  in  gallons  per  minute,  is  known,  then 
the  other  three  can  be  determined  directly  from  the  graphs 
without  further  calculation. 

EXPLANATION. — The  highest  point  on  the  brake-horse-power  graph 
(Fig.  145)  is  about  61  h.p.  This  indicates  that  a  60-h.p.  motor  would 
be  suitable  to  drive  the  pump  at  any  load  without  danger  of  motor- 
overload.  It  is  also  evident  that  the  maximum  efficiency  is  about  73 
per  cent.,  and  that  when  operating  at  this  maximum  efficiency,  about 
1,800  gal.  per  min.  will  be  delivered  against  a  90-ft.  head.  When  oper- 
ating under  these  conditions,  the  power  required  to  drive  the  pump  is 


126 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


about  59  h.p.  Electric  motors  are  usually  designed  to  operate  at  their 
maximum  efficiency  at  the  rated  full  load.  A  60-h.p.  motor  would, 
therefore,  when  driving  the  pump  against  a  90-ft.  head,  be  operating  at 
about  its  maximum  efficiency.  The  maximum  overall  efficiency  would 
be  obtained  with  the  pump  direct  connected  to  a  60-h.p.  motor,  when 
delivering  1,800  gal.  per  minute  against  a  90-ft.  head. 

NOTE. — A  pump,  having  a  head  graph  similar  to  that  of  Fig.  145, 
has  what  is  known  as  a  rising  characteristic.  That  is,  beginning  at 
shut-off,  the  head  developed  increases  up  to  a  certain  point  (about 
600  gal.  per  min.  in  this  pump)  and  then  decreases.  A  slightly  rising 
characteristic  is  usually  desirable.  Note  that  after  this  600-gal.-per-min. 
point  is  passed,  that  the  horse-power  input  is  increased,  and  that  its 
efficiency  increases  up  to  a  certain  point,  and  then  decreases.  Study 
this  graph  of  Fig.  145  to  obtain  a  further  understanding  of  the  relations 
between  head,  efficiency,  horse  power,  and  discharge,  in  a  centrifugal 
pump  which  is  operating  at  a  constant  speed. 

141.  Graphs  Showing  Typical  Relations  Between  Head, 
Volume,  R.P.M.,  And  Efficiency,  In  Good  Commercial 
Centrifugal  Pumps  are  reproduced  in  Figs.  146,  147,  148 


20 


40 


60  60          100          120 

Per  Cent  Of  Rated  Volume 


140 


160 


160' 


FIG.  146. — Relation  Between  Head,  Volume,  R.P.M.,  And  Efficiency,  In  Good  Com- 
mercial Centrifugal  Pumps  Which  Operate  Above  1,800  R.P.M.  Against  Heads  Greater 
Than  50  Ft. 

and  149.  (Marks'  MECHANICAL  ENGINEERS'  HANDBOOK). 
There  is  no  sharp  division  line  between  high  head  and  low 
head.  Low  head  is  in  these  graphs,  assumed  to  mean  less 
than  50  ft.  High  head  is  assumed  to  mean  above  50  ft. 
Low  speed  is  up  to  600  r.p.m.;  moderate  speed,  from  600  to 
1,800  r.p.m.;  high  speed  above  1,800  r.p.m.  These  graphs  do 
not  show  the  performance  of  any  individual  pump,  but  are 


SEC.  141]         CENTRIFUGAL  AND  ROTARY  PUMPS 


127 


the  averages  of  data  obtained  from  a  large  number  of  good 
commercial  pumps,  and  show  what  may  be  reasonably  ex- 
pected of  the  average  pump.  These  curves  are  particularly 


f/lode-rate  speed 

And 
High  Head 


40  60          80  joo          IZO 

p&r  Cent.  Of  Rated  Volume- 


140 


160 


180 


FIG.  147.— Relation  Between  Head,  Volume,  R.P.M.,  And  Efficiency,  In  Good  Com- 
mercial Centrifugal  Pumps  Which  Operate  Between  600  And  1,800  R.P.M.  Against 
Heads  Greater  Than  50  Ft. 

applicable  to  large-capacity  pumps,  as  in  the  smaller  pumps 
efficiency  is  likely  to  be  sacrificed  to  decrease  the  first  cost. 


.'—Per  Centage  Of  RaHol  R.P.M. 


20 


40  60          80  100          110 

Per  Cent,  Of  Rated   Volume 


140 


160         160 


Fia.  148. — Relation  Between  Head,  Volume,  R.P.M.,  And  Efficiency,  In  Good 
Centrifugal  Pumps  Which  Operate  Betwen  600  And  1,800  R.P.M.  Against  Heads  Less 
Than  50  Ft. 

EXAMPLE. — A  centrifugal  pump  has  a  normal  rating  of  8,500  gal.  per 
min.  when  operating  against  an  80-ft.  head  at  1,700  r.p.m.  About  what 
efficiency  should  be  expected  when  the  pump  is  operating  at  its  normal 
rating?  If  the  speed  is  increased  5  per  cent.,  what  discharge  rate  should 
be  expected  if  the  head  pumped  against  remains  constant,  and  what 


128 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


should  be  the  expected  resulting  efficiency?  SOLUTION. — Since  a  speed 
of  1,700  r.p.m.  and  a  head  of  80  ft.  would  classify  this  pump  as  moderate 
speed  and  high  head,  refer  to  the  graphs  of  Fig.  147.  The  100-per-cent. 
r.p.m.  graph,  the  100-per-cent.  head  graph,  and  the  100-per-cent.  volume 
graph  intersect  at  point  A.  It  is  found  that  an  efficiency  of  about  80 
per  cent,  should  be  expected  when  the  pump  is  operating  at  its  normal 
rated  load.  The  105-per-cent.  r.p.m.  graph  intersects  the  100-per-cent. 
rated  load  graph  at  point  B}  which  shows  that  about  112  per  cent,  dis- 
charge rate  and  about  74  per  cent,  efficiency  may  be  expected  with  a 
5-per-cent.  speed  increase. 


zo 


40  60  60  100          IZO 

Per  Cent.  Of  Rated  Volume, 


140 


160 


180 


Fio.  149. — Relation  Between  Head,  Volume,  R.P.M.,  And  Efficiency  In  Good  Cen- 
trifugal Pumps  Which  Operate  Below  600  R.P.M.  Against  Heads  Less  Than  50  Ft. 

142.  The  Effect  Of  Changing  The  Conditions  Under 
Which  An  Actual  Centrifugal  Pump  Operates  will  now  be 
considered.  While  the  effect  of  changed  operating  conditions 
will  always  be  determined  primarily  by  the  theoretical  prin- 
ciples discussed  in  Sees.  118  to  121,  these  theoretical  laws 
cannot,  usually  without  modification,  for  reasons  already 
suggested,  be  applied  to  an  actual  pump.  In  practice,  the 
most  feasible  method  of  ascertaining  the  consequence  of 
altered  conditions  is  a  graphic  one.  For  this  graphic  method, 
a  chart  (Fig.  145),  on  which  are  shown  the  characteristic 
graphs  for  the  pump  under  consideration,  must  be  employed. 

NOTE. — How  To  ASCERTAIN,  FROM  THE  CHARACTERISTIC  GRAPH,  THE 
EFFECT  OF  CHANGING  EITHER  THE  HEAD  OR  THE  DISCHARGE,  AND 
THE  CORRESPONDING  CHANGE  IN  EFFICIENCY  AND  HORSE-POWER  IN- 
PUT, AT  A  CONSTANT  SPEED,  has  already  been  explained  in  Sec.  140. 
A  method  of  obtaining  the  characteristic  graphs  at  any  desired  speed  (within 
reasonable  limits)  from  the  graphs  for  a  given  speed  will  be  presented  in 


SEC.  143]        CENTRIFUGAL  AND  ROTARY  PUMPS 


129 


the  following  Sec.  Then,  after  having  determined  the  characteristic 
graphs  at  any  desired  speed,  the  head,  rate  of  discharge,  efficiency,  and 
brake  horse  power,  can  be  ascertained  at  this  desired  speed.  The 
charts  for  any  pump  at  the  rated  speed  may  be  obtained  by  test,  or, 
usually,  from  its  manufacturer  by  giving  him  a  detailed  description  (all 
name  plate  data  and  serial  number)  of  the  pump  under  consideration. 

143.  A  Change  In  The  Impeller  Speed  Of  A  Centrifugal 
Pump  will  (Sees.  118  to  121)  affect  the  quantity  of  water 
delivered,  the  head  produced,  and  the  horse-power  input. 
The  formulas  by  which  these  variations  are  computed,  for  a 


Discharge 


1100         feOO         1000         ___ 
In    Gallons    Per   Minute 


Fia.  150.  —  Illustrating  Method  Of  Determining  Centrifugal  Pump  Characteristics  At 
Any  Desired  Speed. 

theoretical  installation  without  water-friction,  are  given  in 
Sees.  118,  119,  120,  and  121.  This  theoretical  condition  of  a 
frictionless  installation  is  very  closely  approximated  in 
practice  where  a  pump  is  delivering  water  to  a  stand-pipe 
through  a  short  length  of  pipe  without  bends.  However,  in 
those  installations  wherein  the  friction  head  (Sec.  9)  is  large, 
these  theoretical  formulas  cannot,  without  modification,  be 
employed.  Having  the  characteristic  graph  of  a  pump  for  a 
given  speed,  the  method  of  obtaining  the  graphs  for  any  other 
speed  may  be  understood  from  a  consideration  of  the  following 
example  : 


130  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

EXAMPLE. — A  1,700-r.p.m.  pump,  which  has  a  head  graph  as  shown  in 
Fig.  150  delivers  1,900  gal.  per  min.  in  a  certain  installation  wherein  the 
static  head  is  52  ft.  and  the  friction  head  due  to  the  piping  is  50  ft. 
What  quantity  of  water  will  be  delivered  and  what  head  will  be  produced, 
provided  the  same  piping  is  used,  if  the  speed  is  decreased  to  1,530  r.p.m.  ? 

SOLUTION. — The  friction  head  varies  approximately  as  the  square  of 
the  volume  of  water  delivered.  Therefore,  at  950  gal.  per  min.,  the  friction 
head  =  (950/1900)2  X  50  =  12.5  ft.  Take  other  values  of  discharge  in 
gallons  per  minute  and  compute  the  corresponding  friction  heads  in  a 
similar  manner.  These  values  of  friction  head  laid  off  vertically  upward 
from  the  "52  ft.  static"  line  (Fig.  150)  on  the  corresponding  discharge- 
rate  lines  result  in  the  friction  head  graph.  Next  select  points  such  as 
A,  B,  C,  D,  and  E,  on  the  1700-r.p.m.  head  graph  (Fig,  150)  and  deter- 
mine the  head  and  discharge  rate  corresponding  to  each  point  selected. 
In  Fig.  150,  point  C  corresponds  to  1,750  gal.  per  min.  pumped  against 
a  107-ft.  head  when  the  pump  is  running  at  1,700  r.p.m.  By  For.  (52), 
the  quantity  of  water  delivered  at  1,530  r.p.m.  =  Vgmi  =  (Nz  X  Vgmi)  -5- 
Ni  =  (1,530  X  1,750)  -r-  1,700  =  1,575  gal.  per  min.  By  For.  (53),  the 
head  produced  at  1,530  r.p.m.  =  LhTZ  =  (Nz  +  NJ*  X  Lhn  =  (1,530  -r- 
1,700)2  X  107  =  86.6  ft.  Thus  point  C",  which  is  a  point  on  the  1,530- 
r.p.m.  head  graph,  has  a  value  of  1,575  gal.  per  min.  at  86.6  ft.  Simi- 
larly, determine  the  values  of  points  A',  B',  D',  E',  etc.,  which  correspond 
to  the  values,  of  A,  B,  D,  E,  etc.,  and  plot  points  A',  B',  C',  D',  E',  etc., 
on  the  chart.  A  smooth  curve  drawn  through  A',  B',  C',  D',  etc.,  results 
in  the  1,530-r.p.m.  head  graph  (Fig.  150).  The  intersection  of  the  1,530 
r.p.m.  head  graph  with  the  friction-head  graph  determines  the  quantity  of 
water  discharged  and  the  head  produced,  which,  in  this  case,  is  about  1,580- 
gal.  per  min.  against  about  a  87-ft.  head. 

Ex  AMPLE  .-At  what  speed  must  the  pump  in  the  preceding  example 
be  driven  to  deliver  1,500  gal.  per  min.? 

SOLUTION. — The  l,500-gal.-per-min.  line  intersects  the  friction-head 
graph  at  point  Y,  and  the  1,530-r.p.m.  head  graph  atX.  Therefore,  the 
speed  required  to  deliver  1,500  gal.  per  min.  =  1,530  —  [( Distance  XY  -5- 
Distance  XB}  X  (1,700  -  1,530)]  =  1,530  -  (0.24  X  170)  =  1,530  - 
40  =  1,490  r.p.m.,  approximately. 

NOTE. — THE  POWER  REQUIRED  To  DRIVE  A  CENTRIFUGAL  PUMP 
AT  ANY  SPEED,  other  than  that  upon  which  the  available  characteristic 
graphs  are  based,  may  be  determined  as  follows:  Suppose  the  chart  is 
provided  for  the  pump  when  running  at  1,700  r.p.m.  (Fig.  150)  and  that 
it  is  desired  to  determine  the  power  required  to  drive  the  pump  at  1,530 
r.p.m.  From  the  available  characteristic  head  graph  construct  the  head 
graph  for  the  desired  speed,  as  explained  above.  When  the  pump  is 
running  at  1,530  r.p.m.  and  operating  under  the  conditions  which  are 
represented  by  point  B',  it  will  have  the  same  efficiency  that  it  has  when 
running  at  1,700  r.p.m.  under  the  conditions  which  are  represented  by 
point  B',  also  the  efficiency  at  C',  D',  E',  etc.,  will  be  the  same  as  that  at 


SEC.  144]         CENTRIFUGAL  AND  ROTARY  PUMPS  131 

C,  D,  E,  etc.,  respectively.  Therefore,  by  projecting  vertically  down- 
ward from  D,  (Fig.  150)  it  is  found  that  the  pump,  when  operating  at 
1,700  r.p.m.  has  an  efficiency  represented  by  S,  of  72  per  cent.  Then 
locate  point  S'  equivalent  to  72  per  cent,  vertically  downward  from  D'. 
Sf  is  one  of  the  points  on  the  efficiency  graph  for  1,530  r.p.m.  Other 
points  are  located  in  a  similar  manner  and  the  1,530-r.p.m.  efficiency 
graph  is  drawn.  From  corresponding  values  of  head,  discharge  rate,  and 
efficiency,  the  brake  horse  power  (power  input  to  motor)  can  be  computed 
by  For.  (60)  and  the  graph  can  then  be  drawn,  as  explained  in  Sec.  139. 

144.  The   Methods   Of  Driving  Centrifugal  Pumps  Are: 

(1)  Belt  or  ropes.  (2)  Direct  connected  to  an  electric  motor. 
(3)  Direct  connected  to  a  steam  or  gasoline  engine.  (4)  Direct 
connected  or  reduction-gear  connected  to  a  steam  turbine.  Each 
will  be  discussed: 

145.  A  Belt  Drive  For  Centrifugal  Pumps  is  better  suited 
to  those  of  small  than  to  those  of  large  capacity.     It  should  be 
employed  only  when  direct-connection  is  infeasible.     When 
it  is  desired  to  use  a  belt  drive,  a  pump  which  has  a  relatively 
low  speed  should  be  selected.     In  general,  the  belt  speed 
should  not  be  permitted  to  exceed  about  4,500  ft.  per  min. 
The  pulley  centers  should  be  located  a  reasonable  distance 
apart,  especially  when  there  is  much  difference  in  the  size 
of  the  driving  pulley  and  the  driven  pulley.     The  tight  side 
of  the  belt  should,  when  possible,  be  underneath. 

NOTE.  —  WHEN  THE  ARC  OF  BELT-CONTACT  Is  APPROXIMATELY  180 
DEGREES,  THE  REQUIRED  WIDTH  OF  A  SINGLE  BELT  To  DRIVE  A  CEN- 
TRIFUGAL pump  may  be  computed  by  the  following  formula: 


(61)  :          Ln  .  2.520  X  JW  (inches) 

Wherein:  Lw  =  width,  in  inches,  of  a  single  belt.  Pbhp  =  maximum 
brake  horse  power  required  to  drive  the  pump.  N  =  revolutions  per 
minute  at  which  the  pump  is  to  operate,  d  =  diameter,  in  inches,  of  the 
pulley  on  the  pump  shaft.  To  obtain  the  required  width  of  a  double 
belt,  multiply  the  result  obtained  from  For.  (61)  by  0.625.  The  pulley 
used  on  the  pump  shaft  should  have  a  face  at  least  2  in.  wider  than  the 
belt. 

146.  The  Direct-Connected  Motor  -Driven  Centrifugal 
Pump  (Fig.  157)  is  one  of  the  most  satisfactory  forms  of 
centrifugal-pump  installations.  The  principal  reasons  for 
this  are:  (1)  Saving  of  floor  space.  (2)  Reduction  of  power- 


132  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

transmission  losses.  Since  the  centrifugal  pump  is  a  relatively 
high-speed  machine,  and  as  high-speed  motors  are  cheaper 
than  low-speed  motors,  a  saving  in  the  first-cost  is  obtained. 
By  the  use  of  a  variable  speed  motor,  various  pumping  condi- 
tions (Sees.  118  to  120)  may  be  satisfied  by  the  same  unit. 

NOTE. — THE  ELECTRIC  MOTOR  As  A  DRIVE  FOR  CENTRIFUGAL  PUMPS 
HAS  DECIDED  ADVANTAGES  in  isolated  installations  or  where  no  facilities 
are  at  hand  for  utilizing  the  heat  available  in  the  exhaust  steam. 

NOTE. — DIRECT-CURRENT  MOTORS  find  application  for  installations 
where  only  direct  current  is  available  or  where  adjustment  of  speed  is 
necessary.  The  direct-current  motor  has  the  further  advantage  in  that 
it  can  be  designed  for  any  definite  speed.  Where  the  voltage  is  constant 
either  shunt-wound  or  the  compound-wound,  direct-current  motor  can 
be  used  with  success.  Where  the  voltage  is  variable,  as  in  some  tem- 
porary installation,  particularly  when  fed  from  an  electric  railway  cir- 
cuit (see  Sec.  173),  a  compound  direct-current  motor  should  be  used. 
It  is  generally  recommended  that,  when  direct-current  motors  are  used, 
the  discharge  gate  valve  be  closed  in  starting.  This  procedure  should 
especially  be  followed  with  shunt-wound  motors.  Proper  ventilation 
must  not  be  overlooked  in  motor-driven  installations. 

NOTE. — MOTOR-DRIVEN  CENTRIFUGAL  PUMPS  ARE  USUALLY  DE- 
SIGNED To  OPERATE  AT  SPEEDS  OF  ABOUT  1,100,  1,200,  1,700,  AND  1,800 
R.P.M.,  since  these  are  the  more  usual  "synchronous"  speeds  of  alter- 
nating-current motors.  The  synchronous  speed  of  any  QQ-cycle  alternating- 
current  motor  =  7,200  -f-  number  of  poles.  The  actual  full-load  induction- 
motor  speed  will  be  about  5  per  cent,  less  than  the  synchronous  speed. 
Direct-current  motors  are  often  designed  to  run  at  these  speeds.  This 
renders  a  pump  which  is  designed  to  operate  at  one  of  the  above  speeds 
suitable  for  either  direct-  or  alternating-current-motor  drive. 

NOTE. — THE  POWER-FACTOR-CORRECTING  ABILITY  OF  THE  SYN- 
CHRONOUS MOTOR  Is  INCREASING  THE  DEMAND  FOR  DIRECT-CONNEC- 
TED, SYNCHRONOUS-MOTOR-DRIVEN  CENTRIFUGAL  PUMPS.  If,  however, 
the  brake  horse  power  at  shut-off  is  greater  than  about  35  per  cent,  of 
the  full-load  brake  horse  power,  difficulty  is  likely  to  be  experienced  in 
getting  the  motor  to  pull  into  synchronism. 

NOTE. — THE  SQUIRREL-CAGE  ALTERNATING-CURRENT  INDUCTION 
MOTOR  Is  WELL  ADAPTED  To  CENTRIFUGAL-PUMP  DRIVES  (R.  A.  Fiske) 
because  of  the  simplicity  of  the  motor  and  its  control.  The  first  cost  is 
generally  less  than  that  of  a  motor  of  the  slip-ring  type.  Due  to  the 
squirrel-cage  motor's  characteristic  of  low  starting  torque,  a  valve 
should,  where  one  of  these  motors  is  used,  be  placed  in  the  discharge  line 
to  minimize  the  load  on  the  pump  during  the  starting  period. 

NOTE. — SLIP-RING  INDUCTION  MOTORS  are  preferable  for  centrifugal 
pumps  of  the  larger  capacities  because  of  their  ability  to  start  smoothly 
against  great  torques  without  taking  excessive  currents  from  the  line. 


SEC.  147]         CENTRIFUGAL  AND  ROTARY  PUMPS  133 

147.  A  Steam  Or  Gasoline  Engine,  Direct  Connected  To  A 
Centrifugal  Pump  constitutes  an  economical  method  of  opera- 
tion.    The    speed   of   ordinary   reciprocating   machinery   is, 
however;     relatively    low.     Consequently    this    method    of 
drive  is  only  suitable  for  the  low-  and  medium-head  pumps. 

148.  The    Direct-    Or    Gear-Connected    Steam-Turbine 
Centrifugal -Pump   Drive   is   rapidly   gaining   in   popularity. 
Since  both  the  steam  turbine  and  the  centrifugal  pump  are 
inherently  high  speed  machines,  they  are  admirably  suited  to 
each  other.     The  steam-turbine-driven  centrifugal   pump  is 
even  more  flexible  as  to  speed  variation  than  is  a  motor-driven 
pump.     By  the  installation  of  suitable  governors,  which  are 
actuated  by  the  pump  discharge-pressure,  control  is  obtained 
whereby  the  turbine  speed  is  automatically  adjusted  so  that 
the  head  produced  by  the  pump  remains  constant  over  a 
range  of  from  ^  to  full  pump-capacity.     The  maximum  eco- 
nomical speed  for  large-capacity  pumps  operating  against  low 
heads  is  usually  lower  than  the  minimum  economical  turbine 
speed.     Hence  in  such  installations  the  pump  is  connected 
to  the   turbine   through  specially-designed   reduction  gears. 
This  enables  both  turbine  and  pump  to  be  driven  at  their  most 
economical   speed.     There   is   but   little    power-transmission 
loss  (about  2  per  cent.)  through  a  reduction  gear  of  the  double 
helical  type. 

NOTE. — THE  STEAM  TURBINE  WHEN  EXHAUSTING  INTO  A  VACUUM 
AFFORDS  A  VERY  ECONOMICAL  DRIVE.  (R.  A.  Fiske.)  In  such  units 
the  turbine  may  exhaust  into  a  condenser  (Div.  9)  serving  one  or  several 
of  the  main  generating  units.  Or  the  exhaust  may  be  used  to  advantage 
in  feed- water  heaters  (see  Div.  7).  Where  economizers  are  installed  and 
where  there  would  otherwise  be  an  excess  of  auxiliary  exhaust,  low-pres- 
sure turbines  could  be  used  as  drivers,  the  low-pressure  steam  being 
derived  from  other  auxiliaries  or  from  the  intermediate  receivers  of 
the  main  engines  or  turbines.  The  turbine  can  also  be  arranged  to  ex- 
haust into  the  intermediate  stages  of  the  main  generating  units.  For 
the  larger  units,  it  may  prove  advantageous  to  use  a  low-level  jet  con- 
denser (Sec.  336)  taking  the  condensing  water  from  the  discharge  side 
of  the  pump  and  returning  the  waste  water  to  the  suction  well. 

149.  A  Flexible  Coupling  Should  Always  Be  Used  To  Direct- 
Connect  A  Centrifugal  Pump  To  Its  Motive  Power. — Usually, 
the  pump  and  the  driving  unit  have  two  main  bearings  each. 


134 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


-Brarss  Brushing 


If  a  rigid  flange-coupling  is  used,  the  driving-unit  shaft  and 
the  pump  shaft  become,  in  effect,  a  solid  continuous  shaft,  and 
it  is  practically  impossible  to  align  four  bearings  so  that  each 
will  function  properly  at  centrifugal-pump  speeds.  A  flexible 
coupling  (Fig.  151)  will  compensate  for  slight  inaccuracies  in 
alignment,  and  also  reduce  vibration.  In  a  gear-connected 

steam-turbine  drive,  a  suit- 
"Pubber  Bu5h'ng  able  flexible  coupling  should 

be  used  on  each  side  of  the 

gear. 

NOTE. — It  is  generally  con- 
ceded that  flexible  couplings 
may  not  prove  entirely  "flex- 
ible" on  high-speed  shafts. 
Therefore,  a  rigid  baseplate 
extending  under  pump,  driving 
unit  and  gears,  should  always 
be  provided  to  maintain  the 
shafts  of  the  two  units  in 
good  alignment,  especially 
where  the  pump  is  driven  at 
speeds  above  1,500  r.p.m. 


FIG.  151. — Flexible  Coupling  For  Direct- 
Connecting  A  Centrifugal  Pump  To  Its 
Driving  Unit. 


150.  The  Advantages  Of  The  Centrifugal  Pump  (R.  A.Fiske, 
THE  CENTRIFUGAL  PUMP,  Power  Plant  Engineering,  February 
15,  1921)  are:  (1)  But  one  moving  part.     (2)  No  valves  or  pistons 
to  be  kept  in  order.     (3)  Uniform  pressure  and  flow  of  water.     (4) 
Freedom  from    shock.     (5)    Compactness.     (6)    Simplicity    of 
design.     (7)   Simple  to  operate  and  repair.     (8)  High  rotative 
speed,  allowing  direct  connection  to  electric  motors  and  steam 
turbines.     (9)  In  case  of  stoppage  of  delivery,  the  pressure  cannot 
build  up  beyond  predetermined  working  limits.     (10)  Low  first 
cost.     (11)  Low  rate  of  depreciation. 

151.  The  Disadvantages  Of  The  Centrifugal  Pump  are:  (1) 
The  rate  of  flow  cannot  be  efficiently  regulated  for  wide  ranges 
in  duty.     (2)  The  efficiency  is  not  usually  as  high  as  the  best 
grade  of  piston  pump.     (3)  Direct  connection  to  low-speed  engines 
cannot  be  made  when  operating  on  high  lifts. 

152.  A  Comparison  Between  Centrifugal  Pumps  And  Recip- 
rocating Pumps  will  explain  the  increasing  demand  for  the 
centrifugal  pump.     The  centrifugal  pump  is,  in  general,  su- 


SEC.  153]        CENTRIFUGAL  AND  ROTARY  PUMPS  135 

perior  to  the  reciprocating  pump  in  simplicity,  reliability, 
ease  of  operation,  durability,  space  occupied,  and  frequently 
in  over-all  efficiency.  It  has  a  more  uniform  discharge  pressure 
than  has  the  displacement  pump,  it  vibrates  less  and  does  not 
require  as  heavy  a  foundation.  Except  for  very  small  ca- 
pacities, the  average  first  cost  of  centrifugal  pumps  is  about 
^3  of  that  of  reciprocating  pumps.  The  centrifugal  pump  is 
capable  of  handling  water  which  contains  gravel,  sand,  and 
if  suitably  designed  even  fair-sized  rocks.  This  is  impossible 
with  the  other  type.  The  inherent  characteristics  of  the 
centrifugal  pump  render  it  unsuitable  for  services  which  re- 
quire a  very  positive  control  of  capacity  and  head.  The 
centrifugal  pump  is  not  well  adapted  to  services  which  require 
a  high  suction-lift  (Sec.  88),  nor  for  pumping  small  quantities 
of  water  against  high  heads. 

NOTE. — THE  CENTRIFUGAL  PUMP  WHEN  DESIGNED  FOR  A  CERTAIN 
CAPACITY  AND  HEAD  CANNOT  BE  USED,  WITHOUT  GREAT  Loss  IN 
EFFICIENCY,  AT  ANY  OTHER  CAPACITY  OR  HEAD.  It  is  not  as  flexible 
in  this  respect  as  is  the  reciprocating  pump,  which  can  be  used  under 
widely  different  conditions  without  any  great  sacrifice  in  efficiency. 

153.  The  More  Common  Services  To  Which  Centrifugal 
Pumps  Are  Applicable  are:  (1)  Sewage  pumping  plants.     (2) 
Dry  dock  pumps.     (3)  Irrigation  and  drainage.     (4)  Condenser 
circulating  pumps.     (5)  Municipal  water  works.     (6)  Hydraulic 
elevators.     (7)     Mine    drainage    and    hydraulic    mining.     (8) 
Fire   pumps.     (9)    Boiler-feed  service.     Pumps  of  the  volute 
type  are  more  generally  used  for  the  four  services  which  are 
first  mentioned,  and  those  of  the  turbine  type  for  the  five 
services  last  mentioned. 

NOTE. — Certain  of  these  services  will  be  discussed  in  following  Sections 
but  the  scope  of  this  book  does  not  permit  a  detailed  discussion  of  all. 

NOTE. — About  the  only  services  for  which  centrifugal  pumps  cannot 
be  applied  are  high-pressure-hydraulic-press  and  deep-well  service. 

154.  The  Centrifugal  Pump  Is  In  Almost  Universal  Use 
For  Circulating  The  Condensing  Water  (Div.  9)  in  modern 
power-plant  installations.     It  is  applicable  to  jet-,  barometric-, 
and  surface-condenser  service.     The  steam-turbine  drive  is 
particularly  applicable  to  barometric-condenser  service  (Sec. 


136  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

339)  wherein  a  higher  head  is  required  at  starting  than  under 
normal  operating  conditions.  The  turbine  can  be  speeded 
up  to  produce  the  desired  initial  head.  Then  when  the 
vacuum  becomes  established,  the  speed  can  be  reduced  to 
that  required  for  normal  operating  conditions.  Where  the 
static  head  varies,  as  it  is  likely  to  do  where  the  water  is 
pumped  from  a  river,  variable  speed  operation  is  especially 
desirable. 

155.  Boiler  Feeding  By  Centrifugal  Pumps  (Div.  6)  com- 
prises one  of  the  most  desirable  applications  for  this  type  of 
pump.     It  is,  however,  not  to  be  recommended  for  small-plant 
service.     The  unit  occupies  but  little  space,  and  requires  only 
a  light  foundation.     It  can  be  started  when  cold  and  put  into 
service  in  a  very  short  time.     The  absence  of  vibration  is  an 
important  feature.     Excessive  vibration  in  a  boiler-feeding 
apparatus  will  open  the  pipe  joints. 

NOTE. — A  series  of  tests  which  were  made  by  the  Terry  Steam  Turbine 
Co.  show  an  average  of  62.4  Ib.  of  steam  per  horse-power  hour  for  the 
steam-turbine-driven  centrifugal  pump  as  compared  with  91.9  Ib.  for 
the  reciprocating  pump.  The  tests  were  made  on  boiler-feed  service  at 
a  discharge  rate  of  300  gal.  per  min.  against  a  200-lb.  total  net  head. 

156.  The    Selection  Of  A  Centrifugal  Pump  For  Boiler- 
Feed  Service  (See  Div.  6)  requires,  primarily,  a  consideration 
of:    (1)     Capacity.     (2)     Discharge    or    boiler    pressure.     (3) 
Location  with  respect  to  the  feed  water.     (4)  Load  factor  of  the 
plant.     A  boiler-feed  pump  must  always  be  designed  for  excess 
capacity  over  that  required  for  the  rated  boiler  horse  power 
which  is  to  be  served  (Sec.  225).     Where  high  peak  loads  are 
carried  for  short  periods,  the  installation  of  duplicate  units  is 
advisable.     One  can  then  be  operated  under  normal  loads, 
and  both  during  the  peak-load  period.     When  no  economizers 
are  used  the  discharge  pressure  for  a  centrifugal  boiler-feed 
pump  should  exceed  the  boiler  pressure  by  about  25  Ib.  per 
sq.  in.     When  economizers  are  used,  the  excess  should  be 
from  35  to  50  Ib.  per  sq.  in. 

157.  Where  A  Centrifugal  Pump  Is  Used  For  Boiler  Feed- 
ing In  Connection  With  A  Feed -Water  Heater,  the  hot  water 
should  flow  to  the  pump  under  a  positive  head.     Any  suction 
pull  which  is  exerted  on  water  will  cause  a  reduction  in  the 


SEC.  158]         CENTRIFUGAL  AND  ROTARY  PUMPS 


137 


absolute  pressure  and  a  consequent  lowering  of  the  boiling 
point  and  the  water  will  tend  to  vaporize.  If  the  total  pressure 
exerted  at  the  pump-suction  nozzle  by  the  weight  of  the  hot- 
water  column  is  insufficient  to  overcome  this  tendency,  vapor 
will  be  formed  in  the  pump,  and  the  pump  will  become  vapor 
bound.  This  will  reduce  the  pump-capacity,  or  it  may  cause 
the  pump  to  entirely  stop  delivering  water.  If  the  tempera- 
ture of  the  water  to  be  pumped  exceeds  120  deg.  fahr.  the 
installation  should  be  so  arranged  that  the  suction  head  (Fig. 
153)  is  positive.  For  temperatures  above  120  deg.  fahr.,  there 


7 


140        ,  ISO  160          ITO  180  190 

Approximate  Temperature  In  Degrees  Fahr 


200 


220 


FIG.  152. — Graph  Showing  Effective  Head  Required  At  Pump-Suction  Inlet  To  Suc- 
cessfully Handle  Hot  Water  With  Centrifugal  Pumps.  (Based  On  Data  In  Alberger 
Pump  And  Condenser  Company's  Catalog.) 

should  be  an  effective  head  on  the  center  of  the  pump  equiva- 
lent to  that  shown  by  the  graph  in  Fig.  152. 

NOTE. — CENTRIFUGAL  PUMPS  WILL  DELIVER  WATER  AT  SOMEWHAT 
HIGHER  TEMPERATURES  under  the  corresponding  heads  on  the  suction 
nozzle  than  those  given  by  the  graph  of  Fig.  152,  but  the  efficiency  will 
thereby  be  materially  decreased. 

158.  The  Data  Which  Should  Be  Furnished  The  Pump 
Manufacturer  When  Requesting  Quotations  (R.  A.  Fiske) 
are:  (1)  Capacity  of  pump — gallons  per  minute.  (2)  Total 


138  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

lift,  including  discharge  and  suction  head  as  well  as  pipe 
friction.  (3)  Variations  in  lift,  both  suction  and  discharge. 
(4)  Type — horizontal  or  vertical.  (5)  Quality  of  liquid, 
fresh  water,  gritty  or  solids  in  suspension.  (6)  Temperature 
of  liquid.  (7)  Specific  gravity  of  liquid.  (8)  Service,  water 
works,  irrigation,  boiler-feed  or  what.  (9)  Motive  power 
to  be  used. 

NOTE. — IN  THE  SELECTION  OF  A  PUMPING  UNIT  IT  Is  ALWAYS  BEST 
To  OBTAIN  BIDS  FROM  SEVERAL  MANUFACTURERS;  study  over  the  bids 
carefully;  tabulate  them  in  so  far  as  the  primary  features  are  concerned; 
make  a  careful  comparison  of  details  as  to  ease  of  dismantling,  method 
of  lubrication,  size  and  construction  of  bearings,  materials  used,  and  the 
general  ruggedness  and  serviceability  of  the  pump  as  a  unit.  Then  pur- 
chase the  unit  which,  on  an  annual  cost  basis  (see  Sec.  366  on  Condensers) 
will  probably  be  the  most  economical. 

NOTE. — PROPERTIES  SOMETIMES  DISREGARDED  WHEN  SELECTING  A 
CENTRIFUGAL  PUMP  are:  (1)  Two  single-stage  pumps  each  capable  of 
delivering  500  gal.  per  min.  at  75  ft.  head,  connected  in  series  will  deliver 
500  gal.  per  min.  at  150  ft.  head.  The  same  pumps  connected  in  parallel 
will  deliver  1,000  gal.  per  min.  at  75  ft.  head.  (2)  A  centrifugal  pump 
never  loses  any  head  that  may  be  received  by  it  at  the  suction  chamber. 
For  example :  if  a  pump  capable  of  delivering  water  at}.  200  ft.  head  receives 
it  at  100  ft.  head  it  will  discharge  the  water  at  300  ft.  head.  (3)  When 
two  or  more  equally  rated  centrifugal  pumps  discharge  into  a  common 
main  their  characteristics  should  be  the  same;  otherwise,  under  certain 
conditions  of  head,  they  may  not  give  equal  delivery  and  there  is  the 
possibility  of  one  pump  cutting  out  altogether. 

NOTE. — IT  OFTEN  HAPPENS  THAT  THE  DRIVING  UNIT  SELECTED  Is 
NOT  MANUFACTURED  BY  THE  PUMP  BUILDER.  The  builder  will  always 
quote  on  any  standard  driving  unit  which  may  be  required.  This  raises 
the  question  as  to  whether  or  not  the  purchaser  can  effect  a  saving  by 
buying  the  driving  unit  direct  from  the  manufacturers  and  having  it 
shipped  to  the  pump  manufacturer  and  assembled  to  the  pump  by  the 
latter,  thus  eliminating  the  carrying  charges  expected  by  the  pump 
builder.  The  answer  to  this  is,  avoid  divided  responsibility:  Have  the 
pump  builder  furnish  the  complete  unit,  and  then  hold  him  alone  re- 
sponsible for  its  effectiveness. 

159.  Proper  Installation  Of  A  Centrifugal  Pump  requires, 
first  of  all,  that  the  foundation  be  built  sufficiently  massive 
and  rigid  to  avoid  excessive  vibration.  Vibration  of  the 
stationary  details  of  a  centrifugal  pump  tends  to  produce 
losses,  due  to  excessive  mechanical  friction  and  to  leakage 


SEC.  160]         CENTRIFUGAL  AND  ROTARY  PUMPS 


139 


through  loosened  joints.  The  pump  should  be  set  as  close  as 
possible  to  the  level  of  the  water  at  the  source  of  supply.  It 
will,  in  every  case,  be  advantageous  if  the  pump  can  be  set 
(Fig.  153)  below  the  level  of  the  suction  supply,  so  that  the 
water  may  flow  to  it  by  gravity.  Where  water  at  a  higher 
temperature  than  about  120  deg.  fahr.  is  to  be  handled,  this 
provision  is  practically  imperative.  The  pump  should  be  so 
located  as  to  avoid  elbows,  bends  or  other  sources  of  friction 
in  the  suction  line.  Where  a  suction-lift  cannot  be  avoided 
altogether,  it  should,  if  possible,  be  less  than  15  ft. 


NOTE. — A  SUCTION-LIFT  OF  MORE 
THAN  15  FT.,  including  the  head  due 
to  friction  in  the  pipe,  should  not  be 
attempted  with  pumps  of  larger  size 
than  8-in.  discharge.  With  pumps  of 
smaller  size,  however,  it  may  be  pos- 


Dlscharge _X|<>^J 

,-Pump  Casing 


Sucf/'on  : 
Reservoir-' 


-SucHon  Nozzlt 


Fio.    153. — Centrifugal    Pump    Set    Below 
Level  Of  Suction  Supply. 


FIG.  154. — A    Right-Hand   Centrifugal 
Pump. 


sible  to  operate  with  satisfactory  results  (should  the  conditions  of 
installation  require)  with  a  total  suction-head  up  to  20  ft. 

160.  The  Direction  Of  Rotation  Of  A  Centrifugal  Pump 

should  be  determined  with  a  view  to  convenience  and  adapta- 
bility in  the  installation  of  the  pump.  The  main  factor  to  be 
considered  is  that  the  rotation  shall  be  such  that  the  suction 
and  discharge  nozzles  will  be  so  disposed  as  to  permit  the  lines 
of  suction  and  discharge  piping  to  be  run  as  nearly  straight 
as  possible. 


NOTE. — A  RIGHT-HAND  CENTRIFUGAL  PUMP  (Fig.  154)  is  one  which, 
when  viewed  by  an  observer  from  the  motive-power  end,  rotates  clock- 


140 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


wise.     A  left-hand  centrifugal  pump  (Fig.  155)  is  one  which,  from  the 
same  viewpoint,  rotates  counter-clockwise. 

NOTE. — CENTRIFUGAL  PUMPS  ARE  CONSTRUCTED  WITH  THE  VOLUTES 
OR  CASINGS  So  ARRANGED  (Fig.   156)  As  To  SECURE  CONSIDERABLE 


-Discharge 


Direct/on 


I-Lef1   Hanoi 


[Fia.  155. — A  Left-Hand  Centrifu- 
gal Pump. 


I- Right    Homo! 

FIG.  156. — Diagram  Showing  Various  Dis- 
charge Positions  Of  Centrifugal-Pump  Dis- 
charge-Nozzles. 


Suction 
Nozzle- 


--Discharge  Nozzle 
Driving  Motor-- 
Shaft  Coup/ing-^ 


DIVERSITY  in  the  directions  of  the  discharge  tangents.  This  renders  it 
possible,  in  most  cases,  to  select  a  construction  which  will  permit  of  a 
straight  run  of  discharge  piping  directly  from  the  pump. 

161.  The  Foundation 
For  A  Centrifugal  Pump 
should  be  made  of  con- 
crete or  brick  (see  the 
author's  MACHINERY 
FOUNDATIONS).  It  should 
be  built  up  to  within  about 
0.75  in.  of  the  level  (Fig. 
157)  at  which  the  pump  is 
to  stand.  This  is  to  pro- 
vide space  for  leveling  and 

FIG.  157.— Section  Of  Foundation  Of  Centrifu-     grouting     the     bed-plate   of 
gal  Pump. 

the  pump. 

NOTE. — THE  FOUNDATION  BOLTS,  FOR  HOLDING  DOWN  A  CENTRIFU- 
GAL PUMP,  should  have  at  least  6  in.  of  thread  at  their  upper  ends. 


SEC.  162]         CENTRIFUGAL  AND  ROTARY  PUMPS  141 

Thimbles  (Fig.  157)  made  of  standard  iron  pipe  should  be  set  in  the  ma- 
sonry so  as  to  form  casings  for  the  foundation  bolts.  The  thimbles  should 
be  large  enough  to  give  about  0.5  in.  clearance  around  the  bolts. 

162.  To  Level   A  Centrifugal  Pump,  after  the  pump  has 
been  set  on  the  foundation  with  the  bolts  projecting  through 
the  holes  in  the  bed-plate,  two  or  more  iron  wedges  (Fig.  157) 
should  be  inserted  under  each  of  the  four  edges  of  the  plate. 
The  pump  may  then  be  wedged  up  to  the  proper  level.     The 
packing  in  the  stuffing-boxes  should  then  be  loosened  and  the 
revolving  parts  turned   by  hand  to  test  their  freedom   of 
movement. 

NOTE. — THE  BED-PLATE  OF  A  CENTRIFUGAL  PUMP  SHOULD  BE 
GROUTED  (Fig.  157)  with  liquid  cement  poured  into  the  space  between 
the  top  of  the  foundation  and  the  bed-plate.  When  the  grouting  has  had 
time  to  thoroughly  set  and  harden,  the  foundation-bolt  nuts  may  be 
tightened  down.  Care  should  be  exercised  not  to  draw  the  nuts  so  tightly 
as  to  spring  the  bed-plate.  This,  however,  is  not  likely  to  occur  if  the 
space  beneath  the  plate  is  completely  filled  with  the  grout. 

163.  The  Suction  Piping  Of  A  Centrifugal  Pump  should  in 
no  case  be  of  smaller  size  than  the  suction  orifice  in  the  pump 
casing.     If    the   suction  lift  exceeds   15  ft.,   however,   it  is 

Discharge  Nozzfe,  Jncreofser          5ucfhn  Plpe(of  Larger  Size  Than         -„ .  . 
njf    /    =S=     .r^+»          Suction  Nozzle  Of  Pump)- ^  Well 


m^^^^^^m^-  ~&g&t&?3F&®SB^^^^&  m- 

Driving/}      ''(<-- \- -#0 '  Or  More- '---* •  —  --  ^'-?'~i 

Belt---  "-Suction  Nozzle  At  Least  3  Ft.-~-J> 


FIG.  158. — Long-Draft  Suction  Piping  For  Centrifugal  Pump.  (The  "  increaser" 
should  be  arranged  as  in  Fig.  162  instead  of  as  here  shown.  Also,  the  suction  pipe 
should  pitch  downward  away  from  the  pump.) 

generally  advisable  to  use  piping  one  size  larger  than  the 
suction  opening.  And  if  the  horizontal  distance  over  which 
the  suction  supply  must  be  conveyed  is  very  great,  say  in 
excess  of  100  ft.  (Fig.  158),  piping  at  least  two  sizes  larger  than 
the  inlet  in  the  casing  should  be  used.  This  is  to  avoid 
excessive  friction.  Where  the  water  must  be  lifted,  the  suc- 
tion pipe  should  extend  (Fig. .  158)  at  least  3  ft.  below  the 
level  of  the  water  in  the  suction- well,  or  other  source  of 


142  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

supply.  This  is  to  prevent  air  being  drawn  into  the  pump. 
Scrupulous  care  should  be  exercised,  in  laying  out  the  piping, 
to  avoid  pockets  for  the  accumulation  of  air.  Where  there  is 
any  suction  lift,  all  portions  of  the  piping  should  pitch  down- 
ward toward  the  source  of  supply.  In  such  cases  no  consider- 
able length  of  the  piping  should  run  horizontally,  and  no 
part  of  it,  of  whatever  length,  should  be  allowed  to  pitch 
downward  toward  the  pump.  If  the  water  flows  to  the  pump 
by  gravity,  or  is  supplied  under  the  pressure  of  a  pumping 
system  as  from  street-mains,  a  gate  valve  should  be  inserted  in 
the  suction  line  close  to  the  pump.  This  is  for  convenience  in 


i  ••  i  \«i  ^ 

"tfthytet  Rubbvr-     'Q. 
fvcec/  Yotlvzs.     | 

FIG.  159. — Foot  Valve  With  Stainer.  Fia.  160. — A    Foot- Valve    Disas- 

sembled.    (Strainer    shown    broken 
from  seat-flange.) 

case  it  might  be  necessary  to  disconnect  the  pump  for  repairs. 
If  the  pump  is  required  to  lift  its  suction  supply,  a  foot-valve 
(Figs.  159  and  160)  should  be  connected  to  the  inlet  orifice  of 
the  suction  pipe.  All  right-angled  turns  in  the  piping  should 
be  made  with  long-radius  bends. 

NOTE. — WHERE  A  NUMBER  OF  CENTRIFUGAL  PUMPS  ARE  REQUIRED 
To  TAKE  THEIR  SUCTION  FROM  A  COMMON  SOURCE  each  pump  should 
(Fig.  161)  have  an  independent  suction  line. 

NOTE. — A  FOOT  VALVE  (Figs.  159  and  160)  is  merely  a  check  valve, 
provided  with  a  strainer,  which  is  arranged  for  attachment  to  the  lower 
end  of  a  suction  pipe. 

NOTE. — IT  Is  INADVISABLE  To  TERMINATE  THE  SUCTION  LINE  IN  AN 
ELBOW  CONNECTED  DIRECTLY  To  THE  SUCTION  NOZZLE  of  the  pump. 
An  elbow  so  connected  causes  a  whirling  motion  of  the  entering  water. 


SEC.  164]         CENTRIFUGAL  AND  ROTARY  PUMPS 


143 


This  tends  to  produce  an  irregular  flow.  Also,  if  the  pump  is  of  the 
double-suction  type  (Fig.  130)  it  tends  to  cause  a  condition  of  unbalance 
between  the  pressures  on  the  two  sides  of  the  double  impeller. 

Where  it  is  necessary  to  insert  an  elbow  at  the  pump  end  of  the  suction 
line,  a  short  length  of  straight  pipe  should  intervene  between  the  elbow 


(Drivt'nyBelf 


Discharge  Nozzles  •  • 

i 

..-Pumps - 


Sucfhn  Pipe 
OfNo.B  Pump. 


Sucfion 

Pipe  Of 

Ml  Pump 


V:"-'.Y;.:V.V  ;'v"-'.":.^:'':r^;*  '^'^Y^-'.*:'- Common'  Source  Of  Supply-, 

FIG.  161. — Centrifugal  Pumps   Drawing  Through   Independent   Suction  Lines   From 

Common  Source. 

and  the  suction  nozzle  of  the  pump.  If  the  suction  line  is  of  larger  size 
than  the  nozzle,  then  a  tapered  reducer  may  be  used.  The  reducer 
should,  however,  be  of  eccentric  form  (Fig.  162)  so  as  to  avoid  the  air 
pocket  which  (Fig.  163)  would  result  if  a  reducer  with  concentric  ends 
were  used. 

{•Discharge  Nozzle 

Auction  Nozzle 


FIG.  162. — Showing  How  Enlarged 
Suction  Line  Should  Be  Connected  To 
Centrifugal  Pump. 


Fro.  163. — Showing  How  Concentric 
Reducer  In  Centrifugal-Pump  Suction 
Line  Forms  An  Air-Pocket. 


164.  The  Suction  Pipe  Of  A  Centrifugal  Pump  For  Drawing 
Water  From  A  Driven  Well  (Fig.  164)  should,  if  the  well  is 
deep  enough,  run  down  inside  the  well-casing  to  a  depth  of 
about  25  ft.  The  annular  space  between  the  suction  pipe  and 
the  casing  should  be  sufficient  to  permit  free  access  of  air  to  the 
surface  of  the  water  in  the  well.  This  space  should  be  left 
uncovered  at  the  top  of  the  casing. 

NOTE. — A  SINGLE  CENTRIFUGAL  PUMP  SHOULD  NOT  BE  REQUIRED 
To  LIFT  WATER  FROM  A  BATTERY  OF  WELLS  OR  SUMPS.  A  separate 


144 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


pump  should  be  connected  to  each  well.  Where  the  suction  piping  of  a 
single  pump  is  connected  to  more  than  one  well,  the  pump  will  tend  to 
draw  the  larger  part  of  its  supply  from  the  well  or  wells  nearest  to  it. 


I-Front  Elevation  Of 
Steel  Supporting  Frame 


I-Side  Elevation  Of 
Steel  Supporting  Frame 


FIG.  164. — Centrifugal  Pump  Drawing  Water  From  A  Driven  Well. 

One  of  the  wells  may  thus  be  pumped  down  to  the  level  of  the  inlets  to 
the  piping.  When  this  occurs,  air  will  enter  the  piping  and  break  the 
suction  from  the  other  wells. 


SEC.  165]         CENTRIFUGAL  AND  ROTARY  PUMPS 


145 


NOTE. — A  BY-PASS  BETWEEN  THE  SUCTION-  AND  THE  DISCHARGE- 
PIPING  OF  A  CENTRIFUGAL  PUMP  (Fig.  165)  may  often  be  useful  as  a 
means  of  preventing  the  pump  from  losing  its  suction,  due  to  low  water. 
This  may  occur  where  the  pump  is  used  for  drawing  water  from  a  sump. 
By  adjustment  of  the  valves  A  and  B  the  water  in  the  sump  can  generally 
be  kept  at  any  desired  level. 

165.  An  Air -Chamber  In  The  Suction  Line  Of  A  Centrifugal 
Pump  (Fig.  166)  may  be  necessary  where  the  water  that  is  to 
be  pumped  is  so  impregnated  with  air  that  the  suction  lift 
cannot  otherwise  be  steadily  maintained.  By  running  a  pipe 
from  the  top  of  the  air-chamber  to  the  vacuum  pump  of  a 


Gate  Vcr/res 


'  •  -Discharge, 
Pip* 


'^-•-Vacuum  Pipe  (Connected  To  A 

Yo/cuum  Susfem  Or  To  An  Air 

Pump) 

Air  Chamber-. ,x 
Discharge  P/pe-^^ 
•'Driving  Motor 
\   Centrifuge//  Pump-, 

\ir-~-1N  (. 


mm 


FIG.   165. — Centrifugal   Pump  With   By-        FIG.  166. — Centrifugal-Pump    Suction 
Pass.  Line  Equipped  With  Air-Chamber. 

condenser  or  of  a  heating  system,  or  to  any  system  of  piping 
in  which  a  vacuum  is  maintained,  the  air  can  be  removed  as 
fast  as  it  separates  from  the  suction  water. 

NOTE. — THE  VACUUM  PIPE  leading  from  the  air-chamber  in  a  cen- 
trifugal-pump suction  line  should  extend  to  a  vertical  height  of  at  least 
34  ft.  above  the  level  of  the  suction  nozzle  of  the  pump.  This  is  to 
insure  that  no  water  will  pass  from  the  suction  line  into  the  vacuum 
system  to  which  the  air-chamber  is  piped. 

166.  The  Discharge  Piping  Of  A  Centrifugal  Pump  should 
be  laid  out  with  a  special  view  to  minimizing  pipe  friction. 
Unnecessary  or  avoidable  frictional  resistance  in  the  piping 
means  a  wasteful  expenditure  of  power  in  driving  the  pump. 
The  piping  should  never  be  of  smaller  size  than  the  discharge 
nozzle  of  the  pump.  But  where  pipe  friction,  due  either  to 
excessive  length  of  the  line  or  to  unavoidable  turns  therein,  is 
10 


146 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


a  considerable  item,  it  is  usually  advisable  to  minimize  it  by 
using  pipe  of  a  larger  size  than  the  purnp  connections. 

NOTE. — THE  FLOW  VELOCITIES  AND  CORRESPONDING  FRICTIONAL 
RESISTANCES  OF  SYSTEMS  OP  DISCHARGE  PIPING  may  be  computed,  and 
adequate  pipe  sizes  selected  by  using  the  values  given  in  Table  14  or  15. 

167.  The  Discharge  Pipe  Of  A  Centrifugal  Pump  Should 
Contain  A  Check-Valve  (Fig.  167),  located  as  closely  as 
possible  to  the  pump.  The  function  of  the  check-valve  is  to 
protect  the  pump-casing  from  breakage,  due  to  waterhammer 

that  might  occur  in  the  dis- 
charge line.  Waterhammer  is 
particularly  liable  to  damage  a 
centrifugal  pump  which  is  un- 
protected by  a  check  valve,  if 
a  float  valve  is  attached  to  the 
suction  line. 


NOTE. — A  GATE  VALVE  SHOULD  BE 
INSTALLED  IN  ADDITION  To  the  check- 
valve  in  the  discharge  line  of  a  cen- 
trifugal pump.  The  check-valve 
should  be  connected  between  the  gate 
valve  and  the  pump  discharge  nozzle. 
The  function  of  the  gate  valve  is  to 
afford  a  means  for  controlling  the  dis- 
charge from  the  pump.  It  also  serves 
to  isolate  the  check-valve  from  the 
discharge  piping  in  the  event  of  repair 
or  inspection  of  the  check-valve  be- 
coming necessary. 


FIG.  167. — Priming  Ejector  For 
Use  With  Pump  Unprovided  With 
Foot  Valve.  (Valve  A  is  first  opened. 
The  steam  valve  B  is  then  opened. 
When  water  begins  to  flow  from  the 
ejector  nozzle,  C,  the  pump  is  primed. 
Valves  A  and  B  may  then  be  closed  and 
the  pump  started.) 


168.  Centrifugal  Pumps  Re- 
quire Priming. — That  is,  the 
casing  of  the  pump  must  be 
filled  with  water  before  the  im- 
peller is  set  in  motion.  Where  the  pump  is  below  the  water 
level  of  the  source  of  supply  (Fig.  153)  it  may  be  primed 
simply  by  opening  the  gate  valve  in  the  suction  pipe.  Where 
a  foot-valve  is  used  (Fig.  168)  and  the  discharge  pipe  is  con- 
nected to  an  overhead  tank  or  reservoir,  a  by-pass  (Fig.  168) 
may  be  connected  between  the  discharge  pipe  and  the  suction 
pipe  to  compensate  for  leakage  through  the  foot-valve  while 


SEC.  169]        CENTRIFUGAL  AND  ROTARY  PUMPS 


147 


the  pump  is  shut  down.     The  casing  may  thus  be  kept  full 
of  water  at  all  times. 


Discharge  P/pe — 
Driving  Belt-  -% 


NOTE.  —  A   CENTRIFUGAL  PUMP  SHOULD   NOT  BE   RUN  WHEN  ITS 

CASING    Is   EMPTY.     Certain   of   the 

...  i    i    ••    .,  Priming  By-  Pass  ...... 

interior  parts  are  lubricated  only  by 

the  water  which  passes  through  the 
pump.  Running  the  pump  dry  is 
ruinous  to  these. 

169.  There  Are  Several 
Methods  Of  Priming  Centrifu- 
gal Pumps.  —  A  vacuum  pump 
may  be  connected  (Fig.  169)  to 
the  opening  in  the  top  of  the 
casing.  Then  with  a  check-valve 
in  the  discharge  pipe  or  with  the 
gate  valve  in  the  discharge  pipe 

,          ,     ,  ,          .  FIG.  168.—  Showing  How  The  Dis- 

closed,  the  air  may  be  exhausted    charge  Pipe  of  A  Centrifugal  Pump 


Founcforthn- 


from    the    casing    so   that    the    Should  Be 

water  will  rise  and  fill  the  casing  through  the  suction  pipe. 

The  same  effect  may  be  produced  by  running  a  pipe  from  the 


..-Check  Vafre 

ch<zr-5p 
Pump 


1  /'     t--P!tcher-5pout 


Purnt. 


Hump 

Pipe  ftunning  To ,'     M     Casing* 
Vacuum  Pump-''      (BE)  v 


FIG.  169.—  Centrifugal  Pump  Ar- 
ranged For  Priming  By  Means  Of  A 
Vacuum  Pump. 


FIG.  170. — Priming-Pump  Mounted 
On  Casing  Of  Centrifugal  Pump. 
(Valve  A  is  first  opened.  Handle,  H, 
is  then  worked  until  water  appears  at 
Spout  B.  Valve,  A,  may  then  be  closed 
and  the  pump  started.) 


top  opening  in  the  casing  to  a  steam-condensing  system,  or  to 
any  system  of  piping  in  which  a  vacuum  exists.     In  some 


148 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


situations  it  may  be  convenient  to  fill  the  pump  casing  directly 
from  the  city  water  mains,  or  from  an  elevated  tank  in  a 
house  or  factory  water  supply  system.  Or  a  hand-pump, 
mounted  either  on  the  pump  casing  (Fig.  170)  or  on  the  wall 
nearby  (Fig.  171)  may  serve  as  a  priming  apparatus.  The 
siphoning  effect  of  a  jet  of  steam,  compressed  air  or  water  is 
also  commonly  utilized  in  the  priming  of  centrifugal  pumps. 

Si-earn  Or 
Compressed  Air 
Supply- P/'pe — 

Ejector-' 


Fia.  171. — Priming-Pump  Mounted  On 
Wall.  (Valve,  A,  is  first  opened. 
Handle,  H,  is  then  worked  until  water 
appears  at  nozzle,  N.  Valve,  A,  may 
then  be  closed  and  the  pump  started.) 


FIG.  172. — Priming  Ejector  For  Use 
With  Pump  Provided  With  Foot  Valve. 
(The  nozzle- valve,  A,  is  first  opened.  The 
steam  valve,  B,  is  then  opened.  When 
water  begins  to  flow  from  the  bleeder,  C, 
the  pump  is  primed.  Valves  A,  B,  and  C, 
may  then  be  closed  and  the  pump  started.) 


The  device  which  embodies  this  principle  is  called  a  priming 
ejector,    Fig.  167. 

170.  Where  No  Foot-Valve  Is  Used  (Fig.  167)  the  priming 
ejector  should  be  so  arranged  that  the  current  of  steam  or 
compressed  air  will  draw  the  air  out  of  the  pump  casing.  The 
water  will  then  rise  through  the  suction  pipe  of  the  pump. 
Where  a  foot- valve  is  used  (Fig.  172)  the  ejector  should  be 


SEC.  171]         CENTRIFUGAL  AND  ROTARY  PUMPS 


149 


so  arranged  that  the  current  of  steam  or  compressed  air  will 
draw  the  water  up  through  the  suction  pipe  of  the  ejector 
and  discharge  it  into  the  pump  casing. 


Counter  Shaft- 


NOTE. — WITH  A  Low  SUCTION  LIFT,  6  ft.  or  less,  a  centrifugal  pump 
can  be  primed,  where  the  discharge  pipe  is  filled  with  water  and  no  foot- 
valve  has  been  provided,  by  first  starting  it  in  motion  and  then  letting 
water  flow  into  the  casing  through  the  discharge  pipe.  But  this  is  a 
very  objectionable  method  and  should  not  be  attempted  where  other 
means  are  available.  When  a  centrifugal  pump  is  primed  in  this  manner, 
the  load  is  suddenly  thrown  on  while  the  apparatus  is  rotating  at  a  high 
speed.  The  shock  thus  produced  may  result  in  injury  to  the  impeller, 
shaft  coupling  or  driving  motor. 

171.  A  Centrifugal  Pump  Should  Be  Started  With  The 
Discharge  Valve  Closed. — This  is  generally  necessary  (Sees. 
168  to  170)  to  facilitate  priming 
of  the  pump.  But  aside  from 
this  consideration,  closure  of  the 
discharge  valve  while  the  pump 
is  being  started  is  advisable  in 
order  that  the  full  load  may  be 
imposed  gradually  (Sec.  146) 
upon  the  driving  motor  or 
mechanism.  When  running 
against  a  closed  discharge  valve, 
a  centrifugal  pump  requires  only 
from  35  to  50  per  cent,  of  the 
power  which  it  consumes  when 
running  at  its  economic  dis- 
charge capacity. 


NOTE. — A  CENTRIFUGAL  PUMP  CAN 
BE  RUN  WITH  THE  DISCHARGE  VALVE 
CLOSED  WITHOUT  GENERATING  AN 
EXCESSIVE  PRESSURE  and,  there- 
fore, without  danger  of  rupturing 
the  casing.  But  it  should  not  be  per- 
mitted to  run  in  this  manner  longer  than  about  20  min.  at  a  time. 
When  running  against  the  closed  discharge  valve  the  propeller  churns 
the  water  in  the  casing.  Churning  of  the  water  develops  heat  therein. 
If  continued  long  enough  it  may  result  in  the  water  becoming  heated 


FIG.  173. — C entrifugal  Pump 
Equipped  With  Branch  Discharge  Pipe 
To  Prevent  Churning  When  Pump  Runs 
Against  Closed  Discharge  Valve. 


150 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


Thrust  Bearing— 

Starting  And 
.-Stopping  Snitch 

\ 


to  a  very  high  temperature.     This  might  be  dangerous,  due  to  the 
tendency  of  the  rotating  parts  to  expand  until  seized  by  the  bearings. 

Where  a  centrifugal  pump  is  driven  from  a  line-shaft,  in  a  factory  or 
mill,  it  may  be  inconvenient  to  shut  down  during  lulls  in  the  demand  for 
a  delivery  of  water  from  the  pump.  But  in  such  cases  a  small  branch 
discharge  pipe  (Fig.  173)  should  be  run  from  the  discharge  nozzle  to  the 
suction  well,  so  that  a  small  quantity  of  water  may  pass  through  the 

pump  and  thus  prevent  churning  and 
heating  when  the  discharge  valve  is 
closed.  Where  an  independent  driv- 
ing apparatus,  as  an  electric  motor 
(Fig.  174)  is  used,  it  should  be  shut 
down  when  a  delivery  of  water  is  not 
desired. 

172.  To  Start  A  Centrifugal 
Pump  the  discharge  valve  should 
first  be  closed  and  the  pump 
primed.  First  turn  the  impeller 
over  a  few  times  by  hand  to  allow 
all  air  to  escape  by  way  of  the 
&>  air  cock  at  the  top  of  the  casing. 
After  the  pump  has  been  fully 
primed  (Sees.  168  to  170)  it  may 
be  started.  The  priming  con- 
nection should  be  closed  as  soon 
as  the  impeller  shaft  begins  to 
turn.  The  discharge  valve 

FIG.  174.— Submersible  Type  Of  Vertical    should    remain    closed    Until    full 
Centrifugal  Pump  Installed  In  Sump.       spee(J      Jg      attained.       It     should 

then  be  opened  slowly  so  that  the  motor  may  pick  up  the  full 
load  gradually.  If  the  pressure  does  not  build  up  as  the  speed 
increases,  the  pump  is  not  thoroughly  primed.  In  this  event  the 
pump  should  be  stopped  and  reprimed.  Finally,  the  bearings 
should  be  examined  to  see  that  the  automatic  oilers  are  working 
properly,  and  the  packing  glands  should  be  adjusted  to  allow  a 
reasonable  leakage  from  the  stuffing-boxes.  Leakage  from  a 
stuffing-box  indicates  that  water  is  being  supplied  to  the 
water-seal  ring  which  is  placed  in  the  stuffing-box  and  which 
prevents  air  from  being  drawn  into  the  casing.  Usually,  the 
nuts  on  the  gland  bolts  can  be  drawn  sufficiently  tight  with  the 
fingers. 


SEC.  173]         CENTRIFUGAL  AND  ROTARY  PUMPS  151 

NOTE. — UNDER  No  CIRCUMSTANCES  SHOULD  A  CENTRIFUGAL  PUMP 
BE  RUN  IN  THE  WRONG  DIRECTION.  The  right  direction  is  generally 
indicated  by  an  arrow  (Figs.  154  and  155)  cast  upon  the  casing.  When 
a  centrifugal  pump  is  run  in  the  wrong  direction,  certain  interior  parts, 
which  are  held  in  place  by  screw  threads,  are  liable  to  unscrew  and 
thereby  wreck  the  pump. 

173.  Electrically-Driven    Centrifugal    Pumps    Should    Be 
Operated  Under  A  Steady  Voltage. — No  attempt  should  be 
made  to  operate  an  ordinary  motor-driven  centrifugal  pump 
with  electric  current  from  any  circuit,   as  a  street-railway 
trolley  circuit,  the  voltage  of  which  fluctuates  widely. 

NOTE. — IF  A  MOTOR-DRIVEN  CENTRIFUGAL  PUMP  Is  DESIGNED  To 
RUN  AT  A  SPEED  CORRESPONDING  To  THE  MOTOR  SPEED  at  the  maxi- 
mum value  of  a  fluctuating  voltage,  it  will  deliver  little  or  no  water  when 
the  voltage  is  low.  On  the  other  hand,  if  it  is  designed  to  give  the  desired 
capacity  with  the  motor  speed  corresponding  to  the  minimum  value  of 
the  voltage,  the  motor  may  be  seriously  overloaded  when  the  voltage 
rises  to  its  maximum  value. 

174.  Centrifugal    Pumps  Are  Not  Difficult  To  Maintain 

in  serviceable  condition.  This  is  due  chiefly  to  the  absence 
of  reciprocating  parts  in  their  structural  details.  The  prin- 
cipal details  to  be  looked  after  are  the  shaft-bearings,  stuffing- 
boxes,  and  wearing-rings. 

NOTE. — BEFORE  A  NEW  CENTRIFUGAL  PUMP  Is  PUT  IN  SERVICE 
the  bearings  should  be  carefully  cleaned  with  kerosene  or  gasoline.  The 
oil-wells  should  then  be  filled  with  a  good  quality  of  mineral  oil,  such  as 
is  especially  prepared  for  motor  bearings.  The  oil  should  be  strained  to 
insure  that  no  gritty  matter  is  mixed  with  it.  The  oil  in  the  wells  should 
be  changed  at  proper  intervals,  perhaps  every  two  weeks  in  the  majority 
of  cases.  At  such  times  the  bearings  should  be  thoroughly  washed  with 
kerosene. 

NOTE. — WHEN  A  CENTRIFUGAL  PUMP  Is  USED  FOR  MOVING  CORRO- 
SIVE LIQUIDS  or  sewage,  the  water  used  in  water-seal  connections  of  the 
stuffing-boxes  should  be  piped  from  some  clear-water  source.  In  such 
cases  the  now  unnecessary  openings  which  would  otherwise  be  in  the 
casings  should  be  plugged. 

NOTE. — THE  STUFFING-BOXES  OF  CENTRIFUGAL  PUMPS  SHOULD  BE 
PACKED  with  loose  braided  cotton  packing  impregnated  with  graphite. 
Ordinary  flax  packing  should  not  be  used,  inasmuch  as  the  friction  be- 
tween this  kind  of  packing  and  a  rotating  rod  is  apt  to  be  excessive. 

NOTE. — IT  Is  GENERALLY  ADVISABLE  To  DRAIN  THE  CASING  OF  A 
CENTRIFUGAL  PUMP  WHEN  THE  PUMP  Is  OUT  OF  SERVICE.  This  is 


152 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  4 


imperative  where  the  pump  is  exposed  to  freezing  temperature.  The 
pump  may  be  drained  by  removing  the  plug  (Fig.  131)  in  the  bottom  of 
the  casing. 

NOTE. — WHERE  A  VERTICAL  CENTRIFUGAL  PUMP  Is  REQUIRED  To 
OPERATE  IN  A  SUBMERGED  POSITION  (Fig.  174)  the  shaft  connection  to 
the  motor  should  be  so  designed  as  to  remove  the  ball  thrust-bearing, 
which  sustains  the  weight  of  the  shaft  and  impeller,  entirely  from  contact 
with  the  liquid  that  is  being  pumped.  Adequate  lubrication  of  the  bear- 
ing cannot  otherwise  be  secured. 

175.  A  Rotary  Pump  (Fig.  175)  is  a  positive-action  dis- 
placement pump.  It  should  not  be  confused  with  the  centri- 
fugal pump.  While  the 
motion  of  both  types  of 
pumps  is  one  of  rotation, 
the  principles  involved  are 
entirely  different. 


Discharge  Nozr 
~0 


Discharge 
Nozzle 


Suction 
Nozzle 


FIG.  175. — Positive  Action  Rotary 
Pump.      (Goulds  Mfg.  Co.) 


Fia.   176. — Showing  Operation  Of  A  Rotary 
Pump. 


176.  The  Action  Of  The  Rotary  Pump  may  be  understood 
by  a  consideration  of  Fig.  176.  Suppose  the  pump  is  fully 
primed,  that  is,  the  casing  and  suction  pipe  are  completely 
filled  with  water.  The  shafts,  S,  are  driven  in  the  direction 
as  shown,  by  a  pair  of  spur  gears  which  are  outside  of  the 
casing,  C,  and  are  not  shown  in  the  illustration.  The  liquid 
is  engaged  by  the  teeth  of  the  lobe  gears,  G,  and  being  thus 
confined  in  the  spaces,  B,  by  the  lobe  gear  teeth  and  the  casing, 
is  carried  upward  by  the  rotation  of  the  gears  to  the  discharge 
outlet,  D.  The  teeth  of  the  lobe  gears  are  so  designed  that  at 
every  instant  they  mesh  with  each  other,  thus  preventing  the 
water  from  returning  to  the  suction  side  between  the  gears. 


SEC.  177]         CENTRIFUGAL  AND  ROTARY  PUMPS  153 

177.  The  Advantages  Of  The  Rotary  Pump  are:  (1)  Low 
first  cost.     (2)  Small  dimensions.     (3)  Ease  with  which  it  may 
be  cleaned. 

178.  The  Disadvantages  Of  The  Rotary  Pump  are:  (1)  Noisy 
in  operation.     (2)  Inefficiency  due  to  the  slip  which  is  caused  by 
the  wear  on  the  lobe  gear  teeth.     (3)  Low  speed,  which  usually 
necessitates  the  use  of  reduction  gears  if  motor  or  steam 
turbine  drive  is  employed. 

179.  The    Services  For  Which  Rotary  Pumps  Are  Most 
Commonly  Used   are:  (1)   Fire  protection.     (2)   Pumping  of 
oils,    chemicals,    cider,    vinegar,    etc.     (3)    Circulating   cooling- 
water  for  gas  engines.     (4)  Circulating  oil  for  pipe-cutting  and 
threading  machines.     (5)  Factories  in  which  food  products  are 
handled  in  liquid  form.     The  feature  which  adapts  the  rotary 
pump  to  most  of  these  services  is  the  ease  with  which  it  can  be 
cleaned,     These   pumps  are  manufactured   in  sizes  ranging 
from  the  small  hand-operated  size  to  those  having  a  capacity 
of  900  to  1,000  gal.  per  min.  against  a  230-ft.  elevation. 

QUESTIONS  ON  DIVISION  4 

1.  Why  was  the  development  of  the  centrifugal  pump  retarded  until  recently? 

2.  What  is  a  centrifugal  pump? 

3.  Explain,  using  a  sketch,  the  theory  of  the  centrifugal  pump. 

4.  Upon  what  law  of  physics  is  the  peripheral  speed  of  a  centrifugal-pump  impeller 
based? 

5.  What  is  the  total  head  pumped  against? 

6.  Upon  what  factors  depend  the  quantity  of  water  which  a  centrifugal  pump  will 
deliver? 

7.  What  theoretical  relations  exist  between  the  speed  of  the  impeller  and  the  quantity 
of  water  delivered?     The  head  produced?     The  required  power? 

8.  What  can  be  said  concerning  the  applications  of  the  turbine  and  volute  pumps? 

9.  How  does  increasing  the  number  of  stages  increase  the  head  produced? 

10.  What  are  the  forces  which  tend  to  unbalance  the  impeller? 

11.  Explain  two  methods  of  counteracting  these  forces. 

12.  How  is  end-thrust  taken  care  of  in  double-suction  pumps? 

13.  Name  some  advantages  and  disadvantages  of  the  open  impeller.     Of  the  enclosed 
impeller. 

14.  What  is  the  chief  disadvantage  of  a  vertical-shaft  centrifugal  pump? 

15.  In  what  ways  are  centrifugal  pumps  classified? 

16.  What  are  the  characteristics  of  a  centrifugal  pump?     How  are  they  obtained? 
Draw  and  explain  a  characteristic  graph  for  a  centrifugal  pump. 

17.  Give  four  methods  of  driving  centrifugal  pumps  and  tell  wherein  each  method 
is  applicable. 

18.  Why  is  a  flexible  coupling  used  to  direct-connect  a  centrifugal  pump  to  its  motive 
power? 

19.  Name  the  more  common  services  to  which  centrifugal  pumps  are  applicable. 

20.  Why  must  a  centrifugal  pump  be  installed  with  its  center-line  below  the  supply- 
water  level  when  handling  hot  water? 


154  STEAM  POWER  PLANT  AUXILIARIES  [Div.  4 

21.  Upon  what  two  factors  in  the  installation  of  a  centrifugal  pump  does  successful 
operation  of  the  pump  mainly  depend? 

22.  What  is  the  highest  suction  lift  generally  advisable  for  centrifugal  pumps? 

23.  What  is  a  right-hand  centrifugal  pump?     A  left-hand  centrifugal  pump? 

24.  What  consideration,  in  any  case,  should  decide  whether  a  right-hand  or  a  left- 
hand  centrifugal  pump  should  be  installed? 

25.  Explain  how  a  centrifugal  pump  should  be  leveled  and  grouted. 

26.  Under  what  circumstances  would  it  be  advisable  to  make  the  suction  pipe  of  a 
centrifugal  pump  two  or  more  sizes  larger  than  the  suction  nozzle? 

27.  What  should  be  the  least  depth  of  submergence  of  a  vertical  suction  pipe?     Why? 

28.  How  should  the  suction  line  of  a  centrifugal  pump  be  valved? 

29.  Why  is  it  inadvisable  to  draw  water  from  a  battery  of  driven  wells  with  but  one 
centrifugal  pump? 

SO.  In  centrifugal-pump  operation,  how  may  trouble  due  to  air-impregnated  suction- 
water  be  avoided? 

31.  What  is  the  principal  consideration  in  determining  the  proper  size  of  discharge 
piping  for  a  centrifugal  pump? 

32.  What  is  the  function  of  a  check- valve  in  the  discharge  line  of  a  centrifugal  pump? 
Of  a  gate  valve  in  the  discharge  line? 

33.  What  is  meant  by  priming  a  centrifugal  pump? 

34.  Why  is  it  detrimental  to  run  a  centrifugal  pump  without  liquid  in  the  casing? 

35.  Explain  the  operation  of  priming  a  centrifugal  pump  with  an  ejector  where  no 
foot-valve  is  used.      Where  a  foot-valve  is  used. 

36.  Why  is  it  detrimental  to  prime  a  centrifugal  pump  while  the  impeller  is  in  motion? 

37.  Why  should  a  centrifugal  pump -be  started  with  the  discharge  valve  closed? 

38.  Why  is  it  generally  objectionable  to  run  a  centrifugal  pump  continuously  with 
the  discharge  valve  closed? 

39.  How  may  a  centrifugal  pump  be  piped  so  that  it  may,  with  safety,  be  run  con- 
tinuously with  the  discharge  valve  closed? 

40.  Explain  the  procedure  of  starting  a  centrifugal  pump. 

41.  Why  is  leakage  from  the  stuffing-boxes  of  a  centrifugal  pump  permissible? 

42.  What  damage  may  result  from  running  a  centrifugal  pump  in  the  wrong  direction? 

43.  Why  is  it  inadvisable  to  operate  a  motor-driven  centrifugal  pump  with  electric 
current  from  a  trolley  circuit? 

44.  Mention  some  precautions  which  should  be  adopted  regarding  the  lubrication 
of  centrifugal-pump  bearings* 

45.  How  may  clear  water  be  obtained  for  sealing  the  stuffing-boxes  of  a  centrifugal 
pump  if  the  pump  is  handling  sewerage? 

46.  What  kind  of  packing  should  be  used  in  the  stuffing-boxes  of  a  centrifugal  pump? 

47.  Explain  the  action  of  a  rotary-pump. 

48.  Name  its  advantages.     Its  disadvantages. 

49.  To  what  services  is  it  applicable? 

PROBLEMS  ON  DIVISION  4 

1.  What  must  be  the  theoretical  peripheral  velocity  of  the  impeller  of  a  centrifugal 
pump  to  deliver  water  against  a  total  head  of  160  ft.? 

2.  If  the  impeller  in  Prob.  1  is  to  be  driven  at  1,710  r.p.m.,  what  should  be  its  diameter? 

3.  A  centrifugal  pump  running  at  1,140  r.p.m.  produces  a  head  of  90  ft.     Assuming 
no  head  to  be  lost  in  friction,  what  head  will  the  pump  produce  at  1,600  r.p.m.? 

4.  If  a  centrifugal  pump  delivers  400  gal.  per  min.  when  running  at  1,450  r.p.m., 
what  will  be  its  capacity  when  driven  at  1,600  r.p.m.? 

5.  A  belt-driven  centrifugal  pump  requires  10  h.p.  to  drive  it  at  900   r.p.m.     What 
should  be  the  width  of  the  belt  if  the  driven  pulley  is  7  in.  in  dia.? 


DIVISION  5 


INJECTORS 

180.  An  Injector  (Fig.  177)  is  a  boiler-feeding  device  con- 
sisting of  a  group  of  nozzles  so  arranged  that  a  jet  of  steam 
expanding  therein  strikes  a  mass  of  water  and  condenses. 
Thereby  it  imparts  its  velocity  and  heat  energy  to  the  feed- 
water  which  gains,  in  this  way,  sufficient  momentum  to  force 
itself  into  the  boiler  against  a  pressure  higher  than  that  of  the 
original  steam. 


Regulating 
..-Valve 


''•-Steam  Supply  Line 
Injector-.^ 


FIG.  177.— Illustrating  The  Principle  Of  The  Injector. 

NOTE. — Ordinary  injectors  can  discharge  against  a  pressure  greater 
than  130  per  cent,  of  the  steam-supply  pressure.  Special  injectors  are 
obtainable  which  will  utilize  exhaust  steam  at  atmospheric  pressure  and 
therewith  pump  water  into  a  boiler  containing  steam  at  80  Ib.  per  sq.  in. 
A  brief  explanation  of  the  principles  involved  will  clarify  this  seeming 
mystery. 

155 


156 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  5 


181.  The  Theory  Of  The  Injector  may  be  explained  thus: 
A  pound  of  steam  is  a  reservoir  of  considerable  energy.     Ex- 
panding, in  a  well-designed  nozzle,  from  150  Ib.  per  sq.  in. 
(gage)  down  to  a  24  in.  vacuum,  20  per  cent,  or  about  J^  of 
its  heat  content  is  changed  into  mechanical  energy  of  motion, 
or  kinetic  energy,  amounting  to  188,000  foot-pounds.     If  all 
of  this  kinetic  energy  could  be  utilized,  it  would  force  500  Ib. 
of  water  back  into  the  boiler.     Over  97  per  cent,  of  it,  however, 
is  changed  back  again  into  heat  when  the  steam  jet,  travelling 
at  the  rate  of  40  mi.   per  min.,  projects  itself  against  the 
slowly  moving  mass  of  water. 

NOTE. — The  impact  of  two  bodies  always  results  in  the  generation  of 
heat  at  the  expense  of  kinetic  energy.  Now,  the  remaining  3  per  cent, 
of  kinetic  energy,  after  the  97  per  cent,  has  been  reconverted  into  heat, 
is  sufficient,  theoretically,  to  force  15  Ib.  of  water  back  into  the  boiler. 
But  pipe  friction  and  other  losses  cut  this  down  to  about  13  Ib.  of  water 
pumped  per  pound  of  steam  consumed  at  150  Ib.  per  sq.  in.  pressure. 
The  remaining  97  per  cent,  of  the  energy  which  was  changed  back  into 
heat  and  the  %  of  the  original  heat  content  of  the  steam,  are  not  lost 
but  are  absorbed  by  the  feed  water  and  returned  to  the  boiler. 

182.  The  Essential  Parts  Of  An  Injector  are  shown  in  Figs. 
177  and  178.     These  are  purposely  drawn  out  of  proportion 

so  that  the  characteristic 
shapes  of  the  nozzles  can 
be  discerned  more  clearly. 


Overflow  Outlet 


EXPLANATION.  —  The  steam 
nozzle,  S,  (Fig.  178)  at  the  left,  is 
so  designed  that  the  steam,  in 
passing  through  it,  loses  pressure 
and  gains  a  tremendous  velocity. 
When  a  pound  of  steam  expands 

P  i  G  .  1  7  8  .-Sectional    View   Of   Elementary     f  rom  boiler  F^SSUre  to  a  partial 

Injector.  vacuum  and  to  the  correspond- 

ing lower  temperature,  it  liber- 
ates heat  which  is  converted  into  kinetic  energy  and  thereby  causes  the 
steam  to  attain  a  very  high  velocity.  For  an  explanation  of  the  conver- 
sion of  heat  energy  into  kinetic  energy,  due  to  expansion  through  a 
nozzle,  see  the  author's  STEAM  TURBINES.  The  combining  tube,  C, 
is  a  cone-shaped  nozzle  in  which  the  swiftly  moving  steam  jet  strikes 
the  water  and  is  condensed.  The  delivery  tube,  D,  is  a  diverging 
nozzle.  It  receives  the  combined  jet  of  water-and-condensed-steam 


SEC.  183] 


INJECTORS 


157 


and  gradually  converts  most  of  the  kinetic  or  velocity  energy  of  the 
jet  into  static  energy  or  pressure.  This  is  needed  to  overcome  the 
head  against  which  the  injector  is  discharging.  Overflows,  H,  are 
slots  or  spill-holes,  usually  located  in  the  combining  tube,  to  permit 
excess  water  or  steam  to  escape  when  starting  up.  The  waste- 
valve,  V,  may  be  a  stop  valve 
but  is  usually  a  lift  or  swing 
check  which  closes  automatic- 
ally in  case  that  a  partial 
vacuum  is  formed  in  the  over- 
flow chamber,  O.  Thus,  V, 
prevents  the  inrush  of  outside 
air  that  would  tend  to  scatter 
the  jet.  The  water  in  the  suc- 
tion chamber,  W,  is  drawn  into 
the  combining  tube  by  the  par- 
tial vacuum  which  is  due  to  the 
continuous  condensation  of  the 
steam  therein. 

183.  Injectors  Are 
Classified  as:  (1)  Lifting. 
(2)  Non-Lifting;  depend- 
ing on  whether  or  not  a 
partial  vacuum  is  created 
in  the  suction  pipe  when 
starting  up.  A  non-lifting 
injector  must  always  be 
placed  below  its  source  of 
feed  water  on  this  account. 
Injectors  that  have  one  set 

f  1         IT    TT1*        iTn\   f  FIG.     179. — The    "Hancock    Inspirator,"  A 

of  nozzles  (L  Fig.  179)  for  Double.Tube  Injector.    Jt  is  claimed  thatit 

lifting    the    Water    and    an-    will,    without    adjustment,    operate    on   steam 

other  set,  ^  for  forcing  it  ^Z£%£t£*££$.3Z 
into  the  boiler  are  called  3  to  4  ft.,  lift  oo-deg.  water  25  ft.,  and  with 
double-lube  injector*  Thosp  45  lb  steam  Presssure  wil1  lift  water  25  ft- 

mJ6C  "    and  elevate  it  112  ft.  above  inspirator.) 

which  accomplish  the  same 
result  with  only  one  set  of 

nozzles  (Fig.  180)  are  called  single-tube  injectors.  If  the  oper- 
ation of  an  injector  automatically  re-establishes  itself  after  an 
interruption  in  steam  or  water  supply,  it  is  said  to  be  re-starting, 
or,  more  usually,  automatic.  But  when  the  injector  must  be 


158 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  5 


•Steam  Nozzle 


manually  re-started,  before  it  will  continue  to  operate,  it  is 
said  to  be  positive.  Automatic  adjustment  for  variations  in 
steam  pressure  or  in  height  of  lift  and  temperature  of  feed 
water  is  a  feature  of  self  adjusting  injectors.  All  double-tube 
injectors,  and  a  special  type  of  single-tube  injector  which  has 

a  moving  combining  tube, 
belong  to  this  class.  The 
Sellers  self-acting  injector 
is  both  self  adjusting  and 
re-starting. 

184.  How  An  Automatic 
Injector  Works  is  indicated 
by  Fig.  180  which  shows  a 
section  through  a  Pen- 
berthy  Automatic  I  n  - 
jector.  This  is  a  single 
tube,  re-starting,  lifting- 
type  injector: 

EXPLANATION. — Steam  enters 
at  the  top,  and,  expanding  in  the 
steam  nozzle,  R,  rushes  through 
the  draft-tube,  S,  carrying  with 
it  enough  entrained  air  to  create 
a  partial  vacuum  in  suction 
chamber,  B.  Unable  to  dis- 
charge against  the  boiler  pres- 
sure, this  steam  escapes  through 
the  large  opening  above  the 
sliding  washer,  T,  and  through  the  overflow  opening,  D,  via  P  and  O 
to  the  atmosphere.  The  partial  vacuum  in  B  has  already  lifted  water 
into  it,  and  this  water  has  condensed  part  of  the  steam.  As  more  and 
more  of  the  steam  condenses,  the  jet  becomes  more  compact  and  finally 
becomes  sufficiently  small  to  pass  through  the  least  diameter  of  the 
combining  tube,  C.  Thence  it  passes  through  the  delivery  tube,  Y, 
and  a  check  valve  (Fig.  187)  to  the  boiler. 

The  swiftly-moving  jet  of  water-and-condensed-steam  creates  a  par- 
tial vacuum  in  tube  C.  This  draws  the  loose  washer,  T,  up  against  its 
seat.  Thereby  is  prevented  any  inrush  of  air  which  would  scatter  the 
jet.  The  closing  of  T  also  prevents  any  loss  of  feed  water  through  it. 
If  the  steam  or  water  supply  becomes  interrupted,  the  jet  is  destroyed 
and  the  vacuum  above  T  is  lost.  This  allows  T  to  drop  down  to  its 
original  position.  Hence,  upon  the  resumption  of  the  steam  or  water 
supply,  the  operation  just  described  is  repeated. 


— -Sue  t!on  Chamber 
—-Draft  Tube 
..--Combining  Tube 


Boiler 


Fio.    180. —  "Penberthy"    Single-Tube    Auto- 
matic Injector. 


SEC.  185] 


INJECTORS 


159 


185.  How  A  Positive  Injector  Works  is  indicated  in  Fig.  181 
which  shows  a  section  view  through  a  Metropolitan  Model  0 
Injector.  This  is  of  the  positive,  double-tube,  lifting  type  and 
is  operated  entirely  by  one  handle. 

EXPLANATION. — When  handle,  H,  is  pulled  back  slightly,  steam  is 
admitted  to  the  lower  lifting  nozzle,  N,  through  the  opening  of  the 
auxiliary  valve,  A,  and  of  the  regulating  value,  R.  The  lifting  nozzles. 
N  and  C,  now  begin  to  operate.  The  excess  steam  escapes  through  the 
intermediate  overflow  valve,  O,  and  thence  to  the  atmosphere,  through 
the  final  overflow  or  waste  valve  F.  As  soon  as  water  is  lifted,  it  will 
reach  the  overflow  through  C.  The  operator  then  pulls  the  handle  back 


•Auxiliary  Valve- A  •  in    4 
Lifting  Nozzle— ''y 
Mixing  Nozzle-'"' 

Suction 


..-Overflow 

'-Intermediate  Overflow  Valvg- 
t--5uct'on 
FIG.   181. —  "Metropolitan"  Model-O,  Double-Tube  Injector. 

gradually,  admitting  steam  into  the  main  nozzle,  M ,  through  the  steam 
valve,  V.  The  action  of  this  steam  in  passing  through  the  remaining 
nozzles  has  already  been  explained.  By  the  time  the  handle  has  been 
pulled  back  as  far  as  it  will  go,  the  injector  is  feeding  into  the  boiler 
through  check  valve,  D,  and  the  link,  L,  has  moved  to  the  left  far  enough 
to  close  the  waste  valve,  F,  by  means  of  the  bell  crank  B  and  the  stem  S. 
Regulating  valve  R  is  used  to  control  the  supply  of  water  to  the  injector. 

186.  The  Advantages  Of  An  Injector  are:  (1)  Simplicity. 
(2)  Compactness.  (3)  Low  first-cost.  (4)  High  temperature  of 
feed-water  delivered.  (5)  Ease  of  operation.  (6)  Low  cost  of 
upkeep  and  repairs.  (7)  High  thermal  efficiency,  about  99  per 
cent,  of  energy  put  into  it  is  utilized.  The  absence  of  any 
moving  parts  is  responsible  for  most  of  these  advantages. 


160  STEAM  POWER  PLANT  AUXILIARIES  [Div.  5 

There  are  practically  no  packing  glands  to  be  renewed  and  no 
parts  to  be  lubricated. 

NOTE. — COLD  FEED-WATER  SETS  UP  STRAINS  THAT  ENDANGER  THE 
STRUCTURAL  STRENGTH  OF  A  BOILER.  Hence,  an  injector  is  of  peculiar 
advantage  on  locomotives  where  the  lack  of  space  and  the  use  of  the 
exhaust  steam  for  stack  draft  prevent  the  installation  of  feed-water 
heaters.  These  same  two  conditions  render  the  injector  peculiarly 
applicable  on  locomotives  because  of  its  compactness  and  because  it  is 
many  times  more  economical  than  the  feed  pump  if  the  exhaust  from 
the  latter  is  not  used  to  heat  the  feed- water.  Used  as  an  emergency  feed, 
an  injector  involves  a  minimum  of  overhead  expense. 

187.  The  Disadvantages  Of  The  Injector  are:  (1)  Inability  to 
handle  water  which  is  very  hot.     (2)  Irregularity  of  operation 
under  extreme  variation  in  steam  pressure,  in  temperature  of 
inlet  water  and  in  quantity  of  water  handled.     (3)  Efficiency 
as  a  pumping  unit  is  extremely  low,  never  over  1  or  2  per  cent ; 
that  is,   when  used  in   ordinary   pumping  service — not  for 
boiler  feeding — an  injector  does  not  compare  at  all  favorably 
with  ordinary  pumps  in  economy.     Few  injectors  can  handle 
water  at  150°F.  and  most  of  them  become  inoperative  at 
much  lower  inlet  water  temperatures.     This  is  the  real  reason 
why  injectors  are  not  extensively  used  in  large  power  plants. 
Such  plants  always  have  an  ample  supply  of  exhaust  steam, 
available  from  the  auxiliaries.     If  this  steam  is  not  used  to 
heat  the  feed-water  it  will  be  wasted. 

NOTE. — A  FEED-WATER  HEATER  PLACED  ON  THE  SUCTION  SIDE  OP 
AN  INJECTOR  WOULD  HEAT  THE  WATER  Too  HOT  FOR  ITS  SUCCESSFUL 
OPERATION.  Placed  on  the  discharge  side,  a  feed-water  heater  would  be 
inefficient  because  the  injector  would  deliver  water  to  it  at  such  a  high 
temperature  that  the  heater  would  not  abstract  much  additional  heat 
from  the  exhaust  steam.  To  heat  feed-water  with  live  steam,  when 
exhaust  steam  is  available,  results  in  poor  economy.  The  irregularity 
of  operation  due  to  variations  mentioned  above  is  not,  in  situations  for 
which  the  injector  is  adapted,  a  serious  drawback  and  necessitates  only 
a  reasonable  amount  of  attention  from  the  operator. 

188.  The  Applications  Of  Injectors  Of  The  Different  Types 

will  now  be  considered :  Whenever  it  is  necessary  or  desirable 
to  locate  the  injector  above  the  source  of  feed,  the  lifting  type 
must  be  used.  This  is  especially  true  in  locomotive  practice 
where  it  is  very  advantageous  to  have  the  injector  where  the 


SEC.  189]  INJECTORS  161 

engineer  can  see  the  overflow  outlet.  The  non-lifting  type  is 
simpler,  cheaper  and  of  special  advantage  where  scale-forming 
feed-water  is  used,  because  it  will  not  clog  up  readily  and  is 
very  easy  to  clean.  Double-tube  injectors  will  handle  hotter 
feed-water  through  higher  lifts  than  will  those  of  the  single- 
tube  type.  Hence  they  are  used  exclusively  on  locomotives 
as  a  main  feeding  device,  and,  extensively,  on  board  ship  and 
in  stationary  power  plants  for  emergency  boiler-feeding.  Re- 
starting injectors  are  used  on  small  boats,  traction  and  logging 
engines,  and  in  small  power  plants.  They  are  of  special  ad- 
vantage for  boats,  road  engines  and  similar  applications 
because  the  sudden  interruption  of  water  supply,  due  to  jar  or 
to  movement  of  the  boat,  will  be  taken  care  of  by  the  "  auto- 
matic" feature.  The  "self  acting"  injector  was  designed  for 
locomotive  use  but  is  applicable  where  any  double-tube  type 
is  necessary.  Injectors  are  often  used  for  testing  and  washing 
boilers,  feeding  compound  into  boiler,  and  similar  services. 
189.  A  Simple  Approximate  Equation  Of  The  Injector,  which 
shows  the  relation  between:  pounds  of  water  pumped  per  pound 
of  steam,  the  initial  temperature  of  the  steam,  and  the  final  tem- 
perature of  the  condensed  steam  is  given  below.  It  is  similar  to 
one  proposed  by  Julian  Smallwood  in  his  MECHANICAL  LABORA- 
TORY METHODS.  In  this  equation  radiation  losses  and  the 
amount  of  heat  which  is  changed  into  work  are  neglected. 
These  two  quantities  amount  to  only  1J^  per  cent,  of  the 
total  heat  energy  involved.  See  derivation  below. 


(62)  W™  =        v  °_          (ib.  water/lb.  steam) 

1  fd  —  i  fi 

Wherein  (see  Fig.  182)  Wsw  =  pounds  of  water  pumped  per 
pound  of  steam,  x  =  quality  or  dryness  of  steam,  expressed 
decimally;  if  steam  contains  1  per  cent,  of  moisture,  then  x  = 
0.99;  a  working  average  value  for  per  cent,  moisture  is  2  per 
cent.,  in  which  case  x  =  0.98.  Hv  =  latent  heat  of  vaporiza- 
tion of  steam  at  the  absolute  pressure,  Pa,  at  which  the  injector 
is  receiving  steam,  as  taken  from  a  steam  table,  in  B  .t.u.  Tfa  = 
temperature  of  the  steam,  at  absolute  pressure  Pa,  in  degrees 
fahrenheit.  T/d  =  final  temperature  of  condensed  steam  = 
temperature  of  feed  water  discharged  into  boiler,  in  degrees 
11 


162  STEAM  POWER  PLANT  AUXILIARIES  [Div.  5 

fahrenheit.     T7/*  =  temperature  of  intake  water  to  injector,  in 
degrees  fahrenheit. 

NOTE.  —  THE  MEASURE  OF  THE  ECONOMY  OF  AN  INJECTOR  is  the 
weight  of  water  pumped  per  pound  of  steam  used.  This  value  may  be  de- 
termined by  applying  For.  (62). 

DERIVATION.  —  When  1  Ib.  of  steam  at  some  absolute  pressure  Pa  lb. 
per  sq.  in.,  is  condensed  and  then  cooled  down  to  a  temperature  of  Tfd 
deg.  fahr.,  it  gives  up  a  quantity  of  heat  =  B.t.u.  =  xHv  +  (T/s  —  Tfd). 
Now,  each  1  lb.  of  water  pumped  into  the  boiler  absorbs  heat  energy 
=  B.t.u.  =  T/d  —  T/i.  Then,  neglecting  the  radiation  losses  and  the 
amount  of  heat  which  is  changed  into  work  (both  of  which  amount  to 
only  13^  per  cent,  of  the  total  heat  energy  involved),  the  following  ap- 
proximate relation  exists  in  the  injector,  because  the  heat  absorbed  by 
the  water  must  just  equal  the  heat  given  up  by  the  steam: 

(63)  Heat  absorbed  by  water  pumped    —  Heat  given  up  by  steam  used. 

(64)  Heat  absorbed  per  1  lb.  of  water  pumped  =  T/d  —  T/i 

(64A)  Heat  given  up  per  1  lb.  of  steam  used  =  xHv  +  (T  '/„  —  Tfd) 

Then,  if  Ww  =  weight  of  water  pumped,  in  pounds,  and  W,  =  weight  of 
steam  used,  in  pounds,  it  follows  from  For.  (63)  that: 


(65)  Ww(7Vd  -  Tfi)  =  Ws(xHv  +  (Tf.  -  Tfd}} 

Now,  transposing: 

W«,      xHv  +  (Tt.  -  Tfd) 
^r~  =  -  w  --  w  — 

W8  1  fd  —  1  fi 

But  if  Wgw  is  taken  to  represent  pounds  of  water  pumped  per  pound  of 
steam  used,  then:  Wsv>  =  Ww/Wg.  Now  substituting  this  W,w  for  its 
equivalent  in  For.  (66),  there  results  For.  (62): 

(67)  W8W  =  xHv  r   ^'gT-  —      db.  water/lb.  steam) 

NOTE.  —  To  determine  the  value  of  Wsw  for  any  injector,  it  is  (assum- 
ing that  the  quality  of  the  supply  steam  is  known,  Author's  PRACTICAL 
HEAT,  Div.  19)  merely  necessary  to  observe  (Fig.  182)  the  intake  and  the 
discharge-water  temperatures  at  the  injector,  observe  the  steam  pres- 
sure, substitute  in  For.  (62)  and  solve. 

EXAMPLE.  —  In  testing  an  injector  (Fig.  182)  the  inlet-water  tem- 
perature was  63  deg.  fahr.,  the  discharge-water  temperature  was  202 
deg.  fahr.,  and  the  steam  pressure,  as  indicated  by  the  gage,  was  105  lb. 
per  sq.  in.  The  moisture  content  in  the  steam  was  2  per  cent.  How 
many  pounds  of  water  was  this  injector  pumping  per  pound  of  steam 
which  it  used? 

SOLUTION.  —  The  quality  of  the  steam  =  x  =  1.00  -  0.02  =  0.98. 
The  latent  heat  of  evaporation  of  steam,  as  taken  from  a  steam  table- 


Sue.  190] 


INJECTORS 


163 


at  105  Ib.  per  sq.  in.  gage  (=  105  +  14.7  =  119.7  Ib.  per  sq.  in.  ab- 
solute) is  877  B.t.u.  The  temperature  of  steam  at  119.7  Ib.  per  sq.  in. 
absolute,  as  taken  from  a  steam  table,  is  341  deg.  fahr.  Now  substitute 
in  For.  (62):  W8W  =  [xH,  +  (Tf.  -  Tfd)]/(Tfd  -  Tfi)  =  [0.98  X  877  + 
(341  -  202)]  +  (202  -  63)  =  [859.5  +  139]  -f-  (139)  =  998.5  -r-  139  = 
7.18  Ib.  of  water  per  Ib.  of  steam. 


FIG.  182. — Injector  Arranged  For  Testing. 

190.  To  Compute  The  Horsepower  Actually  Delivered  By 
An  Injector,  apply  For.  (24).     The  amount  of  water  which  the 
injector  is  handling,  may  be  determined  by  weighing  the  water 
before  it  is  pumped. 

191.  The  Performance  Of  An  Injector  Is  Influenced  By 
The  Following  Important  Factors :     (1)  Temperature  of  inlet 
water.     (2)  Height  of  suction  lift.     (3)  Steam  pressure.     The 
action  of  an  injector  depends  upon  the  condensation  of  the 
steam  jet  by  the  incoming  water.     If  this  water  is  too  warm, 
the  injector  will  not  start.     This  limit  is  called  the  over- 
flowing  temperature.     After  the   injector   has   started,   it   is 
possible  to  operate  with  an  intake  water  of  a  higher  tempera- 


164 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  5 


ture,  up  to  a  certain  limit    called   the   breaking  or   limiting 
temperature. 

NOTE. — Fig.   183  shows  how  these  two  temperatures  vary  with  the 
steam  pressure.     Fig.  184  shows  how  variations  in  the  feed- water  tem- 


Sellers  Lifting  Injector  | 
Foot  Lit t)T 


100  125  ISO  115  200  225  250  215  300  35 
Steam  Pressure  in  Lb.  per  So).  Inch 

FIG.    183. — Limiting   And    Overflowing   Temperatures.     (This   figure   was   taken 
from  page  135  of  Kneass'  PRACTICE  AND  THEORY  OF  THE  INJECTOR.) 

peratures  affect  delivery  temperature  of  feed-water,  capacity  of  injector 
and  pounds  of  water  pumped  per  pound  of  steam.  The  height  of  suc- 
tion lift  affects  the  capacity  of  an  injector  as  shown  in  Fig.  185  taken 
from  8  tests  of  a  "Penberthy"  Size  D  Automatic  Injector  operating  at 
80  Ib.  per  sq.  in.  steam  pressure  and  taking  feed-water  at  74  deg.  fahr. 


Height  of  Lift=lFt  (Data  Given  m 
Kneass'  Practise  cmd  Theory  of  the 
Iru'ector) 


75        80       85 
t       W  oi  t  e  r 


90        95       100      105      110       115 
Temperature 


FIG.  184.— Test  Results  Of  A  "Sellers"  No.  8  Self- Ad  justing  Injector. 

Fig.  186  shows  the  variation  in  capacity  of  the  same  Penberthy  Injector 
operating  under  different  steam  pressure  but  with  the  height  of  lift  and 
inlet  water  temperature  constant  at  4  feet  and  74  deg.  respectively. 

NOTE. — THE  REASON  WHY  THE  WATER  PUMPED  PER  POUND  OF 
STEAM  DECREASES  WITH  AN  INCREASE  IN  STEAM  PRESSURE  (Fig.  184) 


SEC.  192] 


INJECTORS 


165 


is  that  the  mechanical  work  done  by  the  injector,  in  pumping  a  given 
weight  of  water  into  the  boiler,  increases  almost  in  proportion  to  the 
steam  pressure  while  the  heat  content  of  the  steam,  and  therefore  its 
ability  to  do  work,  increases  but  slightly.  Between  100  to  200  Ib.  per 
sq.  in.  pressure,  the  heat  content  of  the  steam  increases  by  less  than  1 
per  cent. 


5  /DU 

^700 
^650 
J  600 

"N 

s 

^ 

•^ 

\ 

\ 

\ 

^500 

'G450 
c 

D-400 

u35C 

f 

Ste 

-7 
m  b 

rxrr>] 

est  Re 
jrthyA 
(5iz< 

'ressure 

ult 
uto 

I.:) 

.80 

s  0 

71dt 

;A 

1C  I 

\ 

ijector 

\ 

bFe'ecfw 

otter  At 

74T. 

)     Z     4     6     8    10    12    14    16    18  20 
Height  Of    Lift  In  Feet 

Fia.  185. — Graph  Of  Test  Results  For 
A  "Penberthy"  Size  D  Automatic  In- 
jector Showing  Relation  Between  Ca- 
pacity And  Height  Of  Lift. 


0    ZO  40  60    80  100  1ZO  MO  160 
Steam  Pressure  Lb.Per  Soj-ln.: 

FIG.  186.— Graph  Of  Test  Results 
Of  A  "Penberthy"  Automatic  In- 
jector Showing  Relation  Between 
Steam  Pressure  And  Capacity. 


192.  The  Selection  Of  An  Injector  requires  a  careful  con- 
sideration of  the  three  factors  discussed  in  Sec.  191.  Select 
an  injector  with  a  capacity  in  gallons  per  hour  that  is  30  per 
cent,  in  excess  of  the  amount  of  water  normally  used.  If  the 
amount  of  water  evaporated  per  hour  is  not  known,  approxi- 
mate values  computed  from  the  following  equations,  taken 
from  Sellers'  RESTARTING  INJECTOR,  may  be  used : 

For    horizontal    or    vertical    tubular  boilers: 

(68)  Gal.  per  hr.  =  ~ 
For  water  tube  boilers: 

(69)  Gal.  per  hr.  = 
For  flue  boilers: 

(70)  Gal.  per  hr.  = 


2.42 


1.17 
Wherein  Abh  =  area  of  boiler  heating  surface,  in  square  feet. 

193.  The  Question  Of  What  Type  Of  Injector  To  Use  For 
Any  Given  Service  has  been  previously  discussed  in  Sec. 
188.  It  is  always  best  to  inform  the  manufacturer  as  to  the 
height  of  lift  and  average  temperature  of  feed  water  and  the 


166 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  5 


maximum,  minimum  and  average  steam  pressure,  as  well  as 
the  required  capacity  of  the  injector.  The  injector  will  not 
operate  at  more  than  its  maximum  or  less  than  its  minimum 
capacity.  Table  194  shows  the  list  prices  and  other  data 
for  automatic  injectors  of  a  well-known  make. 

194.   Table    Showing    Capacities,  Pipe    Connections  And 
Approximate  Weight  Of  Injectors. 


Capacity  gal.  per  hr., 

Manu- 

Approxi- 

Pipe 

1  to  3  ft.  lift,  60  to  110 

Shipping 

facturer's 

mate 

connec- 

Ib. per  sq.  in., 

weight, 

size, 

price, 

tion, 

steam  pressure  • 

boxed, 

doci  cm  «i  "f  inn 

j     11 

• 

lh 

Ut/olglltl  tlUIl 

doiicirs 

in. 

Maximum 

Minimum 

ID. 

O 

15.00 

y± 

60 

35 

2.4 

00 

16.00 

H 

80 

45 

2.5 

A 

18.00 

X 

135 

70 

3.5 

AA 

20.00 

y2 

180 

100 

3.5 

B 

25.00 

H 

260 

140 

5.5 

BB 

30.00 

H 

360 

180 

5.5 

C 

40.00 

i 

475 

250 

8,0 

CC 

45.00 

i 

600 

325 

8.0 

D 

55.00 

1H 

800 

425 

12.0 

DD 

60.00 

1H 

1,000 

525 

12.0 

E 

75.00 

iy2 

1,400 

740 

25.0 

EE 

90.00 

iy2 

1,900 

850 

25.0 

F 

110.00 

2 

2,400 

1,275 

37.9 

FF 

125.00 

2 

3,000 

1,600 

39.0 

G 

150.00 

2^ 

3,600 

1,875 

75.0 

GG 

200.00 

2K 

4,200 

2,150 

75.0 

195.  In  Installating  Injectors  the  typical  piping  scheme 
shown  in  Fig.  187  may  be  followed.  .  The  size  of  pipe  to  use 
for  injectors  of  one  make  can  be  found  in  Table  194  under 
PIPE  CONNECTIONS.  The  steam,  suction  and  discharge  pipes 
are  all  of  the  same  size  except  that,  in  the  case  of  a  suction  lift 
exceeding  10  ft.  or  of  a  long  length  of  suction  line,  a  pipe  one 
or  two  sizes  larger  should  be  used  therefor. 

EXPLANATION. — In  Fig.  187,  the  steam  line  should  be  tapped  into  the 
highest  part  of  the  boiler  and  lagged  all  the  way  to  the  injector,  if  possible. 


SEC.  195] 


INJECTORS 


167 


C  is  a  globe-valve.     The  discharge  line  should  follow  as  near  a  straight 
line  as  possible  to  the  boiler-feed  inlet  and  should  be  securely  fastened 


FIG.  187. — Piping  Of  An  Injector. 

throughout  its  entire  length.     A  check-valve,  E,  must  be  installed  as 

shown.     A  is  a  globe  stop  valve  which  can  be  used  to  cut  off  the  boiler 

pressure  from  the  check-valve  so  it  may  be  opened  for  repair.     The 

overflow  is  usually  piped  as  shown.     It  is,  usually, 

best  not  to  discharge  the  overflow  into  the  hot-well 

or  feed  supply  as  it  may  cause  the  suction  water 

to   become   too   hot  to  be  lifted.     The  overflow 

line  must  always  be  open  at  G  to  the  atmosphere. 

The  funnel,  F,  may  be  an  ordinary  one  of  sheet 

metal  or  one   of  the  special  " non-splash"  types 

(Fig.  188)  on  the  market. 

The  suction  line  must  be  absolutely  air  tight  and 
as  free  from  elbows  and  bends  as  possible.  The 
globe  angle  valve  B  takes  the  place  of  one  elbow. 
The  strainer,  S,  should  not  have  any  opening  in 
it  as  large  as  the  steam  nozzle  in  the  injector  and  c  ^xf ,;  I88'7~ 

.  .     Splash"  Funnel   For  In- 

should  have  a  combined  area  of  opening  several   jector  Overflow  Pipe, 
times  as  great  as  the  suction  pipe  itself.    Figs.  182, 

189  and  190  show  commercial  strainers.     The  distance,  h,  should  always 
be  below  20  feet  and  much  less  than  that  if  possible.     Injectors  are  on 


168 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  5 


the  market  that  will  lift  25  feet.     But  high  lifts  reduce  the  capacity  of  an 
injector  as  well  as  its  ability  to  handle  hot  water.     Further  they  make 


FIQ.  189. — Hose-Connection  Strainer  For        FIG.   190. — Pipe-Connection  Strainer  For 
Injector  Suction-Pipe.  Injector  Suction-Pipe. 


• y  -; •*.  o  ; 


FIG.  191. — Injector  Fed  From  Overhead  Tank. 


starting  difficult  and  operation  impossible  when  there  is  even  a  small 
leak  in  the  suction  line. 


SEC.  196]  INJECTORS  169 

If  the  injector  is  fed  from  an  overhead  tank  (Fig.  191)  or  from  city 
supply  under  pressure,  it  is  advisable  to  insert  an  additional  valve,  D, 
(Fig.  187)  which  can  be  permanently  set  so  as  to  throttle  the  pressure 
down  to  the  desired  limit.  Then,  the  valve,  B,  is  used  only  for  opening 
and  closing  the  feed  line.  All  injectors  should  be  braced,  especially 
those  which  are  operated  by  handles.  After  installing  the  piping,  it 
should  all  be  blown  out  with  steam  before  connecting  up  the  injector. 

196.  In  Operating  Injectors,  the  procedure  is  as  follows: 
To  start  an  automatic  injector  be  sure  that  valve  A  (see  Fig.  187) 
has  been  left  open.     Open  slowly  steam  valve,  C,  wide,  now 
open  suction  valve,  B,  wide.     Then  throttle  it  down  until 
there  is  no  discharge  from  the  overflow.     If  the  suction  valve 
is  wide  open  and  steam  still  escapes  from  the  overflow,  it  will 
then  be  necessary  to  throttle  the  steam-supply  valve.     If  the 
discharge  from  the  overflow  is  cool  water,  then  the  suction 
valve  must  be  throttled.     If  there  are  no  unusual  changes  in 
conditions,  the  suction  valve  B  can  be  adjusted  to  give  proper 
supply  and  then  be  permitted  to  so  remain.     An  injector  like 
that  shown  in  Fig.  181  is  operated  entirely  by  one  lever  as 
described  in  Sec.  185. 

197.  Injector  Troubles  And  Their  Correction  are  discussed 
below.     The  more  important  ones  are  listed.     The  correction 
of  other  difficulties  can,  usually,  be  effected  through  a  con- 
sideration of  the  information  given  here : 

IF  AN  INJECTOR  FAILS  To  LIFT  WATER,  it  may  be  due  to  the  following 
causes:  (1)  Leak  in  suction  line.  (2)  Water  too  hot.  (3)  Steam  pressure 
too  low  for  the  lift.  (4)  Suction  strainer  clogged.  (5)  Wet  steam.  (6) 
Nozzles  of  injector  clogged  up  or  covered  with  scale.  (7)  Waste  or  overflow 
valve  stuck  or  leaking.  (8)  End  of  suction  line  not  below  water.  (9) 
Suction  hose  collapsed  by  partial  vacuum.  To  test  for  leaks  in  suction 
line,  screw  a  cap  on  the  end  of  line  in  place  of  the  strainer.  Then  wedge 
the  waste  valve  shut  with  a  piece  of  wood.  When  steam  is  turned  on, 
the  leaks  will  be  detected  easily.  Steam  is  liable  to  be  wet  unless  taken 
from  the  top  of  the  boiler  and  led  directly  to  the  injector.  If  nozzles 
are  clogged  with  scale  they  can  be  removed  and  cleaned.  Coatings  of 
lime  can  be  removed  by  soaking  the  nozzle  several  hours  in  a  solution  of 
ten  parts  water  and  one  part  muriatic  acid. 

IF  AN  INJECTOR  LIFTS  WATER  BUT  DOES  NOT  DELIVER  To  THE 
BOILER  the  trouble  may  be  due  to  (1),  (3),  (5),  (6)  and  (7)  of  the  above 
and  may  also  be  caused  by:  (10)  Faulty  boiler  check  valve.  (11)  Obstruc- 
tion somewhere  in  delivery  pipe.  In  case  of  the  latter  two  difficulties, 
close  valve  A  and  examine  the  check  valve.  If  it  is  lifting  properly 


170  STEAM  POWER  PLANT  AUXILIARIES  [Div.  5 

leave  the  cap  off  and  take  out  the  disk.  Then  start  the  injector.  If  a 
full  stream  of  water  shoots  out  of  the  check  valve,  then  there  is  an  ob- 
struction between  it  and  the  boiler  (most  probably  inside  at  the  opening 
of  the  feed  pipe). 

IF  THE  INJECTOR  STARTS  BUT  BREAKS,  the  trouble  may  be  due  to 
(1),  (3),  (6),  (11),  and  also  to  (12)  An  improper  adjustment  of  the  water 
supply.  If  water  at  the  overflow  is  hot  then  the  supply  is  inadequate 
and  should  be  increased  by  opening  valve  B  wider.  If  it  is  cold  then  the 
supply  should  be  throttled  by  means  of  valve  B. 

WHEN  STEAM  APPEARS  AT  THE  OVERFLOW  the  fault  may  be  (2)  or 
(4)  or  (13)  Too-high  steam  pressure  for  the  lift.  In  this  case  throttle  down 
the  valve  C  until  the  overflow  discharge  ceases.  Every  user  of  injectors 
should  preserve  a  set  of  directions  for  removal  of  injector  parts  and  should 
have  available  spare  nozzles  for  repairs.  Directions  are  gladly  furnished 
by  the  manufacturers. 

QUESTIONS  ON  DIVISION  5 

1.  Explain  how  it  is  that  an  exhaust  steam  injector  can  pump  water  into  a  boiler 
against  the  boiler  pressure. 

2.  Name  four  important  parts  of  an  injector,  giving  functions  of  each. 

3.  Distinguish  between  an  automatic  and  a  positive  injector. 

4.  What  is  a  self  adjusting  injector  and  why  are  all  double-tube  injectors  of  this  class? 

5.  Name  and  explain  six  advantages  of  injectors  over  feed  pumps. 

6.  Why  are  injectors  seldom  used  in  large  stationary  plants? 

7.  Why  are  injectors  always  used  on  locomotives? 

8.  Explain  effect  of:  (1)  Steam  pressure.     (2)  Height  of  lift.     (3)  Temperature  of  inlet 
water  upon  the  capacity  of  an  injector. 

9.  Give  eight  general  rules  which  should  be  followed  in  installing  injector  piping. 

10.  Give  eight  possible  causes  for  an  injector's  inability  to  lift  water  and  state  the 
correction  for  each. 

PROBLEMS  ON  DIVISION  5 

1.  The  following  data  were  observed  during  an  injector  test:  Temperature  of  inlet 
water,  60  deg.  fahr.     Temperature  of  discharge  water,  200  deg.  fahr.     Steam  pressure, 
100  Ib.   per  sq.  in.,  gage.     Moisture  in  steam,  2J>£  per  cent.     Find  value  of  W«u>  or 
pounds  of  water  pumped  per  pound  of  steam? 

2.  Assume  all  data  in  Prob.  1  except  temperature  of  discharge  water.     Find  what 
this  temperature  will  be  if  W.«,  =  10? 

3.  A  water-tube  boiler  has  a  heating  surface  of  500  square  feet.     What  size  injector, 
as  given  in  Table  194,  should  be  used  to  handle  the  feed  water? 

4.  What  size  steam,  suction,  and  delivery  pipes  should  be  used  in  Prob.  3  if  the 
height  of  lift  is  8  ft.?     If  it  is  15  ft.?     If  it  is  20  ft.? 


DIVISION  6 


BOILER-FEEDING  APPARATUS 

198.  Apparatus    For    Feeding    Water   To    Steam    Boilers 

includes  devices  of  three  principal  types:  (1)  Injectors, 
(2)  Pumps,  (3)  Return  traps  or  gravity  apparatus.  Injectors 
for  boiler-feed  service  in  stationary  power  plants  are  usually 
installed  only  as  stand-by  or  emergency  equipment.  Under 
certain  conditions,  however,  they  may  show  an  economic 
advantage  over  other  forms  of  apparatus.  Pumps  are  the 
most  important  boiler-feeding  devices.  Direct-acting  steam 
pumps  (Div.  2),  crank-action  pumps  variously  arranged  and 
driven  (Div.  3)  and  centrifugal  pumps  (Div.  4)  are  all  used  for 
boiler-feeding  as  will  be  explained  herein.  A  few  years  ago 
the  direct-acting  steam 


FeecfP/pe  To  Boilers--'* 


pump  was  the  most 
widely  used  variety  of 
boiler-feed  pump  and  is 
still  a  very  common  va- 
riety. The  use  of  gravity 
boiler-feeding  apparatus 
is  limited  largely  to  small 
steam  heating  and  non- 
condensing  power  install- 
ations. 

NOTE. — The  general  rules 
for  piping  the  principal  de- 
vices which  are  used  in  boiler- 
feeding  are  similar  to  those  for 
steam  piping.  (See  Div.  11 

for  types  of  joints,  specifications  and  allowable  pressures.)  There  are 
usually  at  least  two  feed  pumps  in  a  stationary  power  plant  and  each 
should  be  connected  to  a  common  header  supplying  all  of  the  boilers.  Ex- 
pansion (see  Div.  11)  in  feed  water  lines  is  not  as  great  on  the  whole  as 
in  steam  lines  but  must  be  allowed  for  nevertheless.  The  pump  inlets 
should  be  connected  so  that  water  may  be  drawn  from  two  or  three  sources 

171 


FIG.  192. — A  Direct-Acting  Steam  Pump  For 
Boiler  Feeding. 


172  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

(Fig.  192)  such  as  hot-well,  feed-water  heater  and  city  water-mains  so 
that  feed  water  of  some  sort  is  always  available  during  repairs  or 
emergencies. 

199.  When  An  Injector  Is  Used  As  A  Pump  For  Raising 
And  Forcing  Water  And  Only  As  A  Pump,  it  is  very  inefficient 
inasmuch  as  it  requires  about  five  times  as  much  steam — or 
coal — as  does  an  ordinary  simplex  or  duplex  steam  pump  to 
do  the  same  work.     Hence,  as  a  device  for  merely  handling 
water  where  boilers  are  not  to  be  fed,  the  injector  is,  on  an 
economic  basis,  entirely  out  of  the  running.     Furthermore, 
there  are  a  number  of  troubles  (Sec.  197)  of  the  injector  that 
further  limit  its  usefulness.     The  injector  cannot,  in  practice, 
effectively   handle   water  at   temperatures   exceeding   about 
100  deg.  F.     This  means  that  it  cannot  be  used  advantage- 
ously with  water  which  has  been  previously  heated  with  the 
feed-water  heater.     Hence,  the  injector  cannot  be  used  at  all 
with  an  open  feed-water  heater.     It  may  be  used  with  a  closed 
heater  installed  between  the  injector  and  the  boiler. 

200.  An  Injector  Will  Not  Start  When  Served  By  A  Steam 
Pressure  Much  Lower  Than  That  For  Which  It  Was  Designed. 
Assuming   that   an   injector   is   started   on   the   pressure  for 
which  it  was  designed,  then  if  the  impressed  pressure  increases 
or  decreases  materially  the  injector  will  cease  to  work.     Nor 
will  it  start   again  automatically  upon  resumption  of    the 
steam  pressure  at  which  it  originally  started  and  for  which 
the  engineer  temporarily  adjusted  it.     To  again  cause  it  to 
pump  water,  the  engineer  must  perform  anew  the  starting  and 
adjusting    process.     Furthermore,    material    change    in    the 
water  pressure  of  the  suction  water  which  is  being  handled 
by  the  injector,  will  cause  it  to  cease  operation.     This  necessi- 
tates a  new  adjustment  and  a  new  start.     Often  when  an 
injector  has  been  working  and  has  become  hot,  if  for  any 
reason  it  stops  or  is  stopped,  it  cannot  be  re-started  until  it 
has  been  cooled  completely  by  sousing  it  with  cold  water. 
Obviously,  all  of  the  above  disadvantages  restrict  the  desirable 
applications  of  the  injector  for  boiler-feed  service.     On  the 
other  hand,  the  simplicity,  small  space  occupied,  absence  of 
moving  parts,  and  low  first-cost  of  the  injector  render  its  use 
desirable  under  certain  conditions. 


SEC.  201] 


BOILER-FEEDING  A  PPA  RA  T  US 


173 


201.  The  Injector  Is  Economical  For  Feeding  Boilers  In 
A  Plant  Not  Equipped  With  Means  Of  Feed-Water  Heat- 
ing.— Under  this  condition  the  injector  (Fig.  193)  acts  as  a  com- 
bined pump  and  pre-heater;  and,  as  such,  is  almost  100  per 
cent,  efficient.  The  conditions  favorable  to  injector  installa- 
tion often  obtain  in  temporary  or  out-of-the-way  plants  where 
the  equipment  must  be  minimized,  and  where  the  saving 
which  would  occur  through  the  installation  of  a  feed-water 
heater  is  more  than  offset  by  its  annual  cost;  see  Sec.  246.  Its 


Delivery  Pipe — ,     ._ 
From  Feed  Pump  \     - 


Connection  To 
-City  Water 
~*  Mct/'n- 


Bhw-Oft  And       1 
Feed  Connection     ' 
To  Boiler-- 


Fio.   193. — An  Inspirator  Type  Of  Injector  Piped  Up  For  Boiler  Feeding. 

feature  of  pre-heating  its  feed  water,  makes  the  injector  addi- 
tionally valuable  where  cold  water  is  to  be  fed  into  the  boiler. 
By  pre-heating,  the  strains  which  cold  water  would  cause  in 
the  boiler  are  avoided.  The  proper  combination,  however, 
of  a  pump  with  a  feed-water  heater  is,  as  a  rule,  more  satisfac- 
tory than  an  injector  for  stationary  power  plants.  Injectors 
are  effectively  employed  on  boilers  for  traction-engines, 
small  saw-mill  engines,  hoisting  and  logging  engines,  and  on 
locomotives. 

202.  The  Relative  Efficiencies  of  Steam  Pumps  and  In- 
jectors As  Boiler-Feeding  Devices  are  given  in  Marks' 
MECHANICAL  ENGINEERS'  HANDBOOK  as  follows:  The  effi- 
ciency of  an  injector  considered  merely  as  a  pump  is  very  low, 


174 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


about  1  to  2  per  cent.  As  a  boiler  feed  pump,  in  which  service 
the  heat  in  the  steam  consumed  is  returned  (see  Sec.  181, 
also  Fig.  195)  to  the  boiler,  the  injector  has  an  individual 
efficiency  of  nearly  100  per  cent.  However,  the  injector  is 
not  ordinarily  the  most  economical  device  for  feeding  a  boiler 
since  it  can  handle  only  cold  or  moderately-warm  water  and 
the  effect  is  equivalent  to  heating  the  feed  water  with  live 
steam.  On  the  other  hand,  a  pump  can  handle  water  which 
has  been  heated  to  a  relatively-high  temperature  by  exhaust 
steam  (from  the  main  or  auxiliary  engines)  which  would  other- 
wise be  wasted.  Injector  steam  consumption  is  about  400 
Ib.  per  water  h.p.  hr.;  a  small  direct-acting  pump  consumes 
100  to  200  Ib. 


flnjector 


Heat  Of  Steam  Above  60'Usecl  By  Engine  H60B.t.u. 
Heat  Of  Steam  Above  60 '  Useol  By  Pump    20B.t.u. 

Total  -  //SO&t.u. 
HeatConsumpfhn-          WOO 
FIG.  194. 


Heat  Of 'Steam  Above  60'Vseot  By  Engine  1160 
•     •      •      Loss  By  Injector  88-86  •      2 


B'HeoitConsumptbn  RelativeToA' 
FIG.  195. 


•0985 


FIG.  194. — Direct-Acting  Feed  Pump,  No  Heater.  (B.t.u.  values  in  this,  and 
the  four  following  illustrations,  are  B.t.u.  per  pound  of  steam  delivered  to  the 
engine.) 

FIG.  195.— Injector,  No  Heater. 


far tOf 
rxfjausf: 


Part  Of 
Exhaust 


HevtOfStevmAboveWUseolByEnyine   ll60B.t.u. 
•    -  '      '     •     -Injector      88  • 

Returned  To  Boiler  140  - 

Net  Heat  Useol          1108  • 
(-HeatConsumptlon  Relative  To&*  $&  =  05J5 

FIG.  196. — Injector    And    Exhaust 
Heater. 


HecttOf  Steam  AboveW'UsectByEnyi'ne  IKO&tu. 

Pimm      20   • 

ReturneclTo       140   • 

Total  Heat  Useol    1040  • 

D-Heat  Consumption  ReMiveToA-$$--a882 

FIG.  197. — Direct-Acting  Feed  Pump 
And  Exhaust  Heater. 


203.  Table  Showing  Relative  Economies  Of  A  Non- 
Condensing  Plant  Using  Boiler -Feeding  Devices  Of  Different 
Types.  (Based  on  data  by  D.  C.  Jacobus,  Kent's  MECHANI- 


SEC.  203] 


BOILER-FEEDING  APPARATUS 


175 


CAL  ENGINEEKS'  POCKETBOOK).  See  Figs.  194,  195,  196,  197, 
and  198.  In  each  case  the  values  are  for  the  same  plant  de- 
livering the  same  power  output  from  its  engine.  The  only 
differences  between  the  cases  are  in  the  boiler-feeding  and  feed- 
water  heating  arrangements. 


Relative 

steam 

Per  cent. 

Refer- 

Equipment 

consump- 

steam 

ence 

tion  from 

saving 

letter 

boilers 

Direct-acting    steam    pump    re- 

Fig. 

<§ 

ceiving  water  at  60  deg.  fahr. 

194. 

fc 

and  forcing  it  directly  into  boiler 

T-J  £ 

at  60  deg.  fahr     

1.000 

0.0 

A 

fl}    +^ 

Sffl 

Injector   receiving   water   at   60 

Fig. 

o 

.£3 

deg.  fahr.,  heating  it  to  146  deg. 

195. 

fahr.  and  forcing  it  directly  into 

P 

boiler  at  tliat  temperature  

0.985 

1.5 

B 

Injector  feeding  water  through  a 

Fig. 

heater  in  which  it  is  heated  from 

196. 

1 

146  deg.  fahr.  to  200  deg.  fahr.  .  . 

0.938 

6.2 

C 

1 

Direct-acting  steam  pump  feed- 

Fig. 

1 

ing  water  through  a  heater  in 

197. 

which  it  is  heated  from  60  deg. 

i 

fahr.  to  200  deg.  fahr  

0.882 

11.8 

D 

0> 

Geared    power    pump    mechani- 

Fig. 

.£> 

cally   driven   by  engine  feeding 

198. 

[£ 

water  through  a  heater  in  which 

it  is  heated  from  60  deg.  fahr.  to 

200  deg.  fahr  

0.868 

13.2 

E 

NOTE. — The  direct-acting  steam  pump  (first  item)  has  a  duty  of 
10,000,000  ft.  Ib.  per  100  Ib.  of  coal  when  used  upon  a  boiler  with  80  Ib. 
per  sq.  in.  gage  pressure.  This  corresponds  to  a  over-all  efficiency  of 
about  1.3  per  cent.  Figs.  194,  195,  196,  197  and  198  show  how  a  set  of 
values  such  as  those  above  may  be  obtained.  One  pound  of  steam  de- 


176 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


livered  to  the  engine  is  taken  as  the  unit.     The  heat  in  both  feed-water 
and  steam  above  the  feed-water  temperature  is  considered. 


flKSBJM  (H&OB.t.u.  For  Engine,  5B.t.u.  For  Pump) 


HeoifOfSfeoimAbove60°UseolByEngineAnolPumpll6SB.t.u. 
'     "  Returned  To  Boiler  !40B.f.u. 

Net  Heat  Used          K)?5B.t.u. 
E^Heo/f  Consumption  Relative  Tofcfijj-*  0.868 

FIG.  198. — Power  Feed  Pump  And  Exhaust  Heater.  (It  is  here  assumed  that 
1,160  B.t.u.  must  be  supplied  per  pound  of  steam  required  to  drive  the  regular 
engine  load  as  in  the  four  preceding  figures;  but,  on  account  of  the  additional 
engine  load  due  to  its  having  to  drive  the  pump,  the  engine  will  now  require  more 
steam  in  the  proportion  of  1,165  to  1,160). 

204.  The  Definitions  Of  The  Pump  Designations  Which 
Have  Been  Adopted  In  This  Division  To  Denote  The  Three 
Different  Types  of  Pumps  are  these:  (A)  Mechanically- 
driven  pump,  (Fig.  208)  or  pump  which  is  mechanically 
driven  from  the  engine.  The  drive  may  be  either  direct  by 


Spring  loiter  Pivot- 
(Drlv/ngr  Motor 
\      loiter^ 


,-Wofter  Relief  Valve 


Rxss 
Inlet*, 


Discharge. 


FIG.  199. — Section  Of  Motor-Driven  Triplex  Single-Acting  Boiler  Feed  Pump. 


a  connecting  rod,  gears  or  belt;  or  indirect  through  a  line  shaft 
or  other  common  forms  of  mechanical  transmission.  (B) 
Motor-driven  pump,  or  pump  operated  by  an  electric  motor 
installed  or  used  specially  for  the  purpose  and  belted  (Fig. 
199),  chain  driven,  geared  (Figs.  200  and  201)  or  direct  con- 
nected (Fig.  202)  to  the  pump.  (C)  Steam  driven  pump  or 


SEC.  204] 


BOILER-FEEDING  APPARATUS 


177 


Drt'y/ny  Moto^-.^ 

Pinion  Or?-, 
Motor  Shaft*, 

Crank-  - . 
{.rank-Shaft, 


Feed  Pipe  To  Boilers..^ 
Gate  Valve-.     ~X 


.__.  Wheel  On      ^-Spur  Wheel  On  Intermediate  Georr- 
tronk-Shaft        Shaft^inhr^  On  Opposite  End  Of 
Gear-Shaft,  Drives  Spur  Wheel  On 
Crank- Shaft) 

FIG.  200. — Driving  Motor  Geared  To  Boiler-Feed  Pump. 


.fkcl-r/c  Motor  Mounted  On  Frame  Of  Pump 
Motor- 


FIG.  201. — View  Of  Crank-End  Of  A  Motor-Driven  Boiler-Feed  Pump. 


Gate 

.priving  Motor 
\    Shaft  Coupling^ 


Boilers 
Multi-Stage Centr't- 
fuaorf  Pump-., 


Suction  Pipe  from  Feed-Water  Heater--' 


FIG.  202. — Centrifugal     Boiler-Feed     Pump     Motor-Driven     Through     Direct     Shaft 

Connection, 
12 


178 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


pump  operated  directly  by  steam  from  the  boiler  and  inde- 
pendently of  the  engine.  In  small  plants  they  will  be  recipro- 
cating pumps  of  either  the  simplex  or  duplex  type  (Fig.  203). 


/Globe  Throttle  Valve. (For  Operation  Of  Pump  With- 

Out  Governor  Control) 
fGfobe  Vctlves(Both  Open 
When  Pump  is  Punning       -Weighted  Lever 
Under  Governor  Control)  /  Feeci  Une  To  ^^ 
governor  to/ve  „•' 


Pipe.  For  Conveying  Feed- 

Line  Pressure  fo  Governor 

Diotphrotgm--^ 


FIG.  203.  —  Direct-Acting  Boiler-Feed  Pump  Equipped  With  Fulton  Governor. 

In  plants  of  over  500  h.p.,  they  may  be  either  reciprocating 
or  centrifugal  pumps  (Fig.  204). 


Multistage  Cen-  feed  'Pipe  Jo  Boilers        Pressure 


trifugal  Boiler 
Feed  Pump-, 


Steam  Turbine-*. 


Governor--^ 


^Suction  Pipe  From  Feed-  Water  Hectter  ^Bchausf  Pipe 


Fia.  204. — Centrifugal  Boiler-Feed  Pump  Driven  By  Steam  Turbine  Through  Direct 

Shaft  Connection. 

205.  The  Possible  Losses  Due  To  Inefficient  Boiler-Feed 
Pumps  Are,  In  The  Average  Plant,  A  Very  Small  Proportion  Of 
The  Total  Losses. — The  total  coal  required,  directly  or  indi- 
rectly, for  the  boiler-feed  pumps  in  a  reasonably-well  designed 


SEC.  206]  BOILER-FEEDING  APPARATUS  179 

and  operated  plant  of  medium  capacity  is  not  liable  to  exceed 
more  than  1  or  2  per  cent,  of  the  total  coal  burned.  In  a  very 
small,  inefficient  plant,  the  proportion  of  the  coal  required  for  the 
boiler-feed  pumps  may  in  exceptional  cases  be  as  great  as  10 
per  cent. 

206.  The  Most  Economical  Type  Of  Boiler -Feed  Pump  And 
Drive  Therefor  Depend  On  Local  Conditions. — Whether  the 
plant  is  condensing  or  non-condensing,  its  horsepower  capacity, 
the  character  of  its  load  and  the  fluctuations  thereof,  the  oppor- 
tunity to  utilize  exhaust  steam,  and  other  special  and  financial 
conditions  may  be  factors. 

207.  The  Actual  Cost  Of  Operation  Of  Any  Boiler-Feed 
Pump  Cannot  Be  Based  Merely  On  The  Efficiency  Of  The 
Pump  Itself.— Nor  can  a  comparison  of  the  actual  operation 
costs  of  boiler-feed  pumps  of  the  different  types  be  based 
merely  on  the  performance  of  the  pump.     The  economic  rela- 
tion of  the  other  components  of  the  plant  wherein  the  pump  is 
to  be  installed  must  be  considered.     Whether  or  not  the  ex- 
haust steam  from  a  steam-driven  feed  pump  can  be  utilized 
for  boiler-feed-water  heating  or  for  building  heating  may  be  a 
determining  factor. 

208.  At   Least   One   Reciprocating  Steam-Driven  Boiler- 
Feed  Pump  Should  Be  Installed  In  Every  Plant. — If  there  is 
only  one  boiler-feed  pump  in  a  plant,  it  should  be  steam  driven 
and  preferably  direct-acting.     The  reason  for  this  is,  as  is 
explained  in  detail  elsewhere  in  this  Div.,  the  inherent  reli- 
ability of  the  direct-acting  steam-driven  pump,  due  to  its  sim- 
plicity and  the  fact  that  there  are  no  links,  except  a  steam  line, 
between  it  and  the  boiler.     Another  advantage  is  that  a  steam- 
driven  feed  pump  can,  if  there  is  steam  in  the  boiler,  be  operated 
whether  or  not  the  engine  is  running. 

209.  A    Motor-Driven    or    a    Power-Driven    Boiler -Feed 
Pump   Is  Always  Most  Economical  In  A  Non-Condensing 
Plant  (if  no  live  steam,  in  addition  to  exhaust  steam,  is  required 
for  building  heating).     The  reason  is  that  a  non-condensing 
engine  of  itself  will  always  (Sec.  240)  furnish  much  more  than 
sufficient  exhaust  steam  to  heat  the  feed  water.     Both  the 
power-driven  and  the  motor-driven  pumps  require  considerably 
less  coal  for  their  operation  than  does  a  steam-driven  pump. 


180 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


Hence,  under  these  conditions,  all  the  heat  in  the  exhaust  steam 
from  a  steam-driven  pump  would  be  wasted.  (It  requires, 
under  ordinary  non-condensing  conditions,  approximately 
14  per  cent,  or  %  of  the  exhaust  steam  from  the  engine  to  heat 
the  feed  water  up  to  212  deg.,  which  is  ordinarily  the  highest 
feasible  feed-water  temperature.  The  remaining  %  or  86 
per  cent,  of  the  exhaust  steam  from  the  engine  is  wasted  into 
the  atmosphere). 

210.  In  Any  Plant  Where  Low-Pressure  Live  Steam  In 
Addition  To  Exhaust  Steam  Is  Required  For  Building  Or  Other 
Heating,  A  Direct-Acting  Steam-Driven  Pump  Will  Ordi- 

Exhctust  To  Atmosphe. 
\  Bore*  Pressure  ya/ve-l- 


-Automottic  Relief  ; 
Valve 

Exhaust  To-1' 
Atmosphere 
—-Condenser 

-Motor-Driven  Air 
Pump 


Steam  Driven  Circulating  Pump' 


FIG.  205.  —  Condensing    Plant    Equipped    With     Motor-Driven    Boiler-Feed     Pump. 
(Cochrane   Heater.) 

narily  Be  Preferable  because  of  its  simplicity,  reliability,  and 
low  first  cost.  Its  steam  consumption  under  these  conditions 
is  of  minor  importance  because  all  of  the  heat  in  its  exhaust 
steam  is  utilized  for  building  or  other  heating.  For  building 
heating,  exhaust  steam  is  nearly  as  effective  as  live  steam. 

211.  The  Use  Of  Some  Non-Condensing  Steam-Driven 
Auxiliaries  Is  Ordinarily  Economical  In  Condensing  Steam 
Power  Plants  (Fig.  205)  .  —  In  condensing  plants  there  is,  as  a 
rule,  no  exhaust  steam  available  from  the  main  engines  for 
feed-water  heating.  When  no  economizer  is  used,  the  auxil- 
iary drives  should  be  so  proportioned  that  there  will  be  just 


SEC.  212] 


BOILER-FEEDING  APPARATUS 


181 


enough  exhaust  steam  from  them  to  heat  the  feed-water  to 
210  deg.  fahr.  This  condition  gives  a  maximum  of  economy 
as  practically  all  of  the  energy  of  the  steam  delivered  to  the 
auxiliaries  is  then  effective  either  as  mechanical  energy  or  heat. 
When  an  economizer  is  used  in  addition  to  an  exhaust  heater, 
some  other  feed-water  heater  discharge  temperature  such  as 
150  deg.  fahr.  may  prove  economical.  Both  motor-driven  and 
steam-driven  pumps  are  often  installed  under  such  conditions. 
Then  the  operator  uses  whichever  pump  or  pumps  the  exhaust 
steam  from  which  will  give  the  feed-water  temperature  from 
the  exhaust  heater  which  has  previously  been  found  to  be  most 
economical. 

212.  An  Automatic  Exhaust  Steam  Heat  Balance  System  (Fig. 
206)  has  been  devised  for  maintaining  a  perfect  balance  between 


-Eleetrk  Control  Wires 
.-St-ctck 


vena*?. 


FIG.  206. — Diagram  Of  Plant  With  Condenser,   Feed-Water  Heater  And  Automatic 
Arrangement  For  Maintaining  Heat  Balance. 

the  exhaust  steam  available  and  that  needed  for  feed-water 
heating.  When  this  system  is  used,  most,  or  all,  of  the  auxil- 
iaries are  motor-driven  and  an  auxiliary  non-condensing- 
engine-generator  unit,  B,  is  provided  to  supply  the  auxiliary 
motors  C  and  F  with  electrical  energy.  Then  this  auxiliary 
generator  is  interconnected  with  the  main  generator  through  a 
motor-generator  set  G,  as  shown  in  Fig.  206.  Under  normal, 


182  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

full-load  conditions,  the  motors  driving  auxiliaries  take  all  of 
their  electrical  energy  from  the  auxiliary  generator,  in  which 
case  it  will  then  produce  just  enough  exhaust  steam  to  heat  the 
feed  water  up  to  212  deg.  But,  if  due  to  change  of  load,  the 
amount  of  exhaust  steam  supplied  by  this  auxiliary  engine- 
generator  unit  becomes  more  than  sufficient  to  heat  the  feed 
water,  then  a  portion  of  the  electrical  energy  which  the  auxil- 
iary motor  drives  take  is  shifted  over  to  the  main  generator  E. 
This  shifting  of  the  electrical  load  is  effected  through  the  motor- 
generator  G. 

EXPLANATION. — This  motor-generator  is  actuated  by  an  electrical- 
contactor  pressure  valve  H  on  the  feed-water  heater  A.  When  there  is 
a  surplus  of  exhaust  steam  in  the  feed-water  heater,  the  valve  contactor 
operates  in  such  a  way  that  the  motor-driven  auxiliaries  take  a  greater 
proportion  of  their  energy  from  the  main  generator.  When  there  is 
insufficient  exhaust  steam  in  the  heater,  the  valve  contactor  operates  in 
the  other  direction,  causing  the  auxiliary  motors  to  take  more  of  their 
energy  from  the  auxiliary  generator.  This  throws  a  greater  load  on  the 
auxiliary  generator  and  results  in  the  production  of  more  exhaust  steam 
by  the  auxiliary  engine-generator  unit.  By  means  of  this  automatic 
arrangement,  a  practically-perfect  heat  balance  always  obtains.  This 
results  in  the  highest  possible  economy  of  operation. 

213.  The  Most  Economical  Boiler-Feeding  Equipment  For 
A   Non-Condensing   Plant   Which   Includes   An   Extensive 
Heating  System  For  Winter  Service  is  (Fig.  207)  a  steam- 
driven  pump  for  operation  during  the  heating  season  and  a 
motor-driven  or  power-driven  pump  for  operation  when  the 
heating  system  is  out  of  service. 

EXPLANATION. — If  the  heating  system  requires  more  steam  in  the 
winter  than  the  main  engine  furnishes  as  exhaust,  some  live  steam  must 
be  drawn  through  a  reducing  valve  for  heating  purposes.  A  steam 
driven  pump  acts  as  a  reducing  valve  and  then  furnishes  power  as  a 
sort  of  by-product.  The  cost  of  the  power,  in  this  case,  for  pumping 
will  be  negligible.  On  the  other  hand,  in  the  summer  when  the  heating 
system  is  not  in  use  there  will  be  a  great  surplus  of  exhaust  steam  from 
the  main  engines  alone.  When  this  is  true  the  motor  driven  or  mechanic- 
ally driven  feed-pump  has  the  economic  advantages  as  shown  in  Sec.  217. 

214.  Mechanical-Drive  For  A  Boiler-Feed  Pump  Is,  Con- 
sidering The   Feed-Pump   Independently,   Generally   More 
Efficient  Than  Electric-Drive  (Figs.  199,  200,  202,  and  208). 


SEC.  214] 


BOILER-FEEDING  APPARATUS 


183 


xhaust  To  Atmosphere-  •  'f<O          =» 


Steam  Pump  For  Bo/ler-Feeof 
,--5errke  In  Winter 


&^££l'£^^ 


FIG.  207.  —  Non-Condensing  Plant  Equipped  With  Steam-Driven  and  Motor-Driven 
Boiler  Feed  Pumps. 


^Countershaft 

rMngr  Pulley 
Driving  Belts 


,-Dr/Yen  Pulley 
,-Connectingr  ffocf 
\  .-Plunger  Roof 


Crotnk- 


FIG.  208. — A   Mechanically-Driven   Boiler  Feed-Pump.     (For   boilers   carrying  steam 
pressure  under  85  pounds  per  sq.  in.) 


184  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

This  is  due  to  the  fact  that,  under  average  conditions,  the  total 
mechanical  losses  between  the  main  engine  and  a  mechanically- 
driven  pump  are  less  than  the  total  mechanical  and  elec- 
trical losses  between  the  main  engine  and  a  motor-driven 
pump.  It  should  be  understood,  however,  that  this  rule 
holds  only  under  favorable  conditions.  If,  in  order  to  drive  a 
feed  pump  mechanically,  it  is  necessary  to  transmit  the  power 
through  long  line  shafts,  through  many  belts,  and  through 
right-angle-turn  belt  transmissions,  then  the  higher  theoretical 
efficiency  of  the  mechanical  drive  will  vanish.  Furthermore, 
the  efficiencies  upon  which  the  following  example  is  based, 
relate  to  mechanical  transmissions  which  are  well  installed  and 
maintained.  Line  shafts  out  of  alignment,  worn  bearings,  and 
similar  conditions  may  result  in  material  decrease  in  efficiency. 
It  should  be  noted,  then,  that  while  the  over-all  efficiency  of 
the  electrical  drive  will  remain  practically  constant  throughout 
the  life  of  the  drive,  because,  the  elements  which  affect  it  are 
the  generator,  electrical  transmission  and  motor  efficiencies,  all 
of  which  are  unaffected  as  time  progresses,  the  efficiency  of  the 
mechanical  drive  may  decrease,  due  to  use. 

EXAMPLE. — Suppose  that  power  is  transmitted  from  the  main  engine 
to  a  mechanically  operated  feed-pump  through  two  successive  belt  con- 
nections— the  first  being  from  the  engine-shaft  to  a  line-shaft,  and  the 
second  from  the  line-shaft  to  the  pinion-shaft  of  the  pump.  Also,  sup- 
pose the  transmission  efficiency  of  the  belting  is  97  per  cent.,  of  the  line 
shaft  96  per  cent.,  and  of  the  pump  gearing  96  per  cent.  The  overall 
efficiency  of  the  mechanical  drive  is,  then,  0.97  X  0.96  X  0.96  =  89.4 
per  cent.  Suppose,  also,  that  power  is  transmitted  from  the  main  engine 
to  a  motor-operated  feed-pump  which  is  working  under  the  same  service 
conditions  as  the  mechanically-driven  pump.  Then,  assuming  a  gener- 
ator efficiency  of  93  per  cent.,  a  transmission  line  or  wiring  efficiency  of 
95  per  cent.,  a  motor  efficiency  of  85  per  cent.,  and  a  pump-gear  efficiency 
of  96  per  cent.,  the  overall  efficiency  of  the  motor-drive  is  0.93  X  0.95  X 
0.85  X  0.96  =  72  per  cent.  The  economical  advantage  of  the  mechan- 
ical over  the  motor  drive  is,  therefore,  represented  by  an  efficiency  differ- 
ence of  89.4  -  72  =  17.4  per  cent. 

215.  Motors  For  Driving  Feed  Pumps  should  be  of  enclosed 
or  semi-enclosed  types  if  the  pumps  are  installed  in  a  dusty 
boiler  room  or  in  any  other  dusty  place.  The  motor  should 
preferably  be  of  the  adjustable-speed  type  so  that  the  water 


SEC:  216]  BOILER-FEEDING  APPARATUS  185 

may  be  pumped  into  the  boiler  at  the  same  rate  as  that  at 
which  it  is  evaporated.  The  rate  of  evaporation  varies  with 
the  load. 

NOTE. — AN  ADJUSTABLE  SPEED  MOTOR  is  one  the  speed  of  which  can 
be  varied  over  a  considerable  range  and  when  once  adjusted  remains 
practically  unaffected  by  the  load.  Examples  are  shunt  wound,  lightly- 
compound-wound  d.c.  motors.  A  varying-speed  motor  is  one  the  speed 
of  which  varies  with  the  load,  such  as  a  d.c.  series  or  heavily-compound- 
wound  motor  or  an  a. -c. wound-rotor  slip-ring  induction  motor.  Since 
adjustable-speed  motors,  capable  of  sufficient  speed  variation  for  efficient 
boiler  feed  service,  are  not  ordinarily  obtainable  in  the  smaller  capacities, 
it  is  necessary  to  use  varying-speed  motors  for  these  small-capacity 
applications. 

216.  If  A  Constant-Speed  Motor  Is  Used  On  A   Boiler 
Feed  Pump  either  the  water  feed  must  be  intermittent,  which 
is  undesirable,  or,  if  the  motor  continues  to  operate  at  constant 
speed,  a  part  of  the  feed  water  must  be  by-passed  through  a 
by-pass  valve.     Where  a  by-pass  valve  is  used,  the  motor  may 
operate  continuously  at  constant  speed  and  little  or  much  of 
the  water  it  pumps  be  admitted  to  the  boiler  by  controlling 
the  by-pass  valve  as  occasion  requires.     This  by-passing  is 
very  uneconomical  because  then  all  the  water  handled  must 
be  pumped  against  boiler  pressure.     Then  the  energy  imparted 
to  the  portion  of  the  water  which  is  not  fed  into  the  boiler  is 
wasted.     This  situation  may  be  practially  corrected  in  the 
larger  plants  by  installing  two  feed  pumps,  each  of  one-half 
the  capacity  necessary  for  total  requirements. 

NOTE. — GEAR  DRIVE  Is  PREFERABLE  To  BELT  DRIVE,  because  of 
high-cost  maintenance.  Feed  pumps  are  frequently  installed  in  out-of- 
the-way  corners  where  it  is  difficult  to  make  prompt  repairs  on  belts. 
The  belt  may  slip  off  or  break  when  such  an  accident  can  be  least  afforded. 

217.  The    Economy   Of   A   Mechanically-Driven   Boiler- 
Feed  Pump  Which  Operates  At  A  Constant  Speed  is  affected 
adversely  where  the   load   on  the   plant   fluctuates  widely. 
This  is  due  to  the  fact  that  the  water  horsepower  output  of 
the  pump  cannot  be  varied  economically  in  response  to  the 
fluctuations  of  the  load. 


186  STEAM  POWER  PLANT  AUXILIARIES  [Diy.  6 

NOTE. — A  mechanically-driven  pump  for  boiler-feeding,  and  its  driver, 
should  (Sec.  231)  be  designed  to  meet  the  maximum  requirements  of  the 
boilers.  The  quantity  of  water  delivered  to  the  boilers  may  be  regu- 
lated (Fig.  209)  in  accordance  with  load  variations,  by  means  of  a  by-pass. 
But  with  this  method  of  control,  a  large  proportion  of  the  power  con- 
sumed by  the  pump  is  wasted.  Although  the  quantity  of  water  passing 
into  the  boilers  through  the  check-valves  in  the  delivery  pipes  may  be 
diminished,  the  pressure  at  the  by-pass  connection  to  the  discharge  pipe 
will  only  be  slightly  less  than  the  boiler  pressure.  The  pump  will,  there- 
fore, still  be  discharging  at  its  full  capacity  against  almost  full  boiler 
pressure.  The  average  economy  of  a  mechanically-driven  boiler-feed 
pump  (considered  as  an  independent  unit),  operating  under  the  most 
extreme  conditions  of  load-fluctuation  which  might  prevail,  would 
nevertheless  be  generally  superior  to  that  of  a  steam-driven  pump  oper- 
ating under  like  conditions. 


Belt-Connection  To  Line-Shcrft--. 
Countershaft- — 


Outs/c/e-Cinte-r- 

Packeof  Plunger 

Pump..., 


Fio.  209. — Mechanically-Driven  Boiler-Feed  Pump  Furnished  With  By-Pass. 

EXAMPLE. — Suppose  a  steam-driven  pump,  which  consumes  175  Ib. 
of  steam  per  h.p.  per  hr.,  develops  4  h.p.  while  feeding  a  fully  loaded  set 
of  boilers.  Then  the  quantity  of  steam  required  to  operate  this  pump 
will  be  175  X  4  =  700  Ib.  per  hr.  Also,  suppose  a  mechanically-driven 
pump,  operated  by  main-engine  power  which  is  transmitted  to  the  pump 
on  a  steam  consumption  of,  say,  30  Ib.  per  hr.,  develops  4  h.p.  while 
feeding  a  similar  set  of  boilers.  Then  the  quantity  of  steam  required 
to  operate  this  pump  will  be  30  X  4  =  120  Ib.  per  hr.  If  the  load  on 
each  set  of  boilers  is  diminished  one-half,  then  the  power  required  to  feed 
them  will  be  4  -r-  2  =  2  h.p.  The  steam  pump  need  then  be  run  at  only 
one-half  its  former  speed.  Hence,  it  will  develop  no  more  than  the 
requisite  2  h.p.  during  the  period  of  half  load.  Its  steam-consumption 
will,  therefore,  be  reduced  to  700  -j-  2  =  350  Ib.  per  hr.  The  mechanic- 


SEC.  218]  BOILER-FEEDING  APPARATUS  187 

ally-driven  pump  will,  however,  maintain  its  original  speed.  It  will, 
therefore,  continue  to  discharge  at  the  rate  of  its  full-load  capacity. 
But  only  one-half  of  the  water  discharged  will  enter  the  boiler.  The 
remaining  half  will  be  by-passed.  All  of  the  water  will,  nevertheless,  be 
discharged  against  the  boiler  pressure.  Hence,  the  pump  will  continue 
to  develop  4  h.p.  during  the  half-load  interval.  Consequently,  it  will 
maintain  its  original  steam  rate  of  120  Ib.  per  hr.  Its  economical 
advantage  over  the  steam  pump  will,  therefore,  be  reduced  from 
[(700  -  120)  -T-  700]  X  100  =  83  per  cent,  under  full-load  conditions  to 
[(350  -120)  -5-  350]  X  100  =  66  per  cent,  under  half-load  conditions. 

218.  The  Saving  Which  May  Be  Effected  By  Substituting 
An  Electrically-Driven  For  A  Steam-Driven  Feed  Pump 

may  be  estimated  as  follows:  The  example  is  based  on  a 
non-condensing  steam  plant,  which  has  a  feed-water  heater, 
and  which  operates  twenty-four  hours  a  day.  It  is  assumed 
that  the  conditions  are  such,  as  is  usually  the  case  in  a  non- 
condensing  plant,  that  the  exhaust  steam  from  the  steam- 
driven  pump,  direct-acting  or  turbo-centrifugal,  cannot  be 
utilized  effectively  for  feed-water  heating  or  otherwise. 
(The  engine  alone  in  a  non-condensing  plant  furnishes  about 
six  or  seven  times  as  much  exhaust  steam  as  can  possibly  be 
reclaimed  for  feed-water  heating.  Hence,  in  such  a  plant  the 
exhaust  steam  from  the  feed  pump  represents  pure  waste.) 

EXAMPLE.  —  Average  load  on  plant,  50  kw.  Assumed  water  rate  for 
this  simple  engine  plant,  50  Ib.  of  steam  per  kw.-hr.  Therefore,  the  steam 
consumption  per  hr.  for  this  50-kw.  plant  =  50  X  50  =  2,500  Ib.  of 
steam  per  hr.  Hence,  Gal.  of  feed  water  required  per  hr.  =  2,500  -s-  8.3  = 
300  gal.  per  hr.  (approx.).  Pressure  against  which  feed  water  is  forced  = 
125  Ib.  per  sq.  in.  Head  =  Ib.  per  sq.  in.  X2.31.  Hence,  for  this  plant, 
head  =  125  X  2.31  =  290  ft.  approximately.  Work  required  per  hour 
to  force  water  into  boiler  =  weight  of  water  per  hr.  X  head  =  2,500  X 
290  =  725,000  ft.-lb.  per  hr. 

Now  the  average  expected  duty  of  an  ordinary  4-in.-stroke  steam- 
driven  boiler-feed  pump  may  be  stated  conservatively  as  11,700,000 
ft.-lb.  per  1,000  Ib.  of  steam.  Or,  in  other  words,  it  may  be  assumed 
safely  that  for  a  steam  boiler-feed  pump  in  this  50-kw.  plant: 


1  1 

No.  of  ft.-lb.  per  Ib.  of  steam  =          QQ'         =  11,700  ft.-lb. 

As  is  evident  from  preceding  statements,  the  pounds  of  steam  con- 
sumed per  hour  in  this  plant  to  develop  the  725,000-ft.-lb.  required  to 


188  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

pump  the  2,500  Ib.  of  water  against  the  125  Ib.  per  sq.  in.  steam  pressure 
is: 

T ,       ,   j.  j  •  ft.-lb.  to  supply  water  per  hr. 

Lb.  of  steam  to  drive  steam  pump  per  hr.  =  - — 

ft.-lb.  per  Ib.  of  steam 

=        '    ^   =  62  Ib.  of  steam  per  hr. 
1 1,  /  UU 

That  is,  a  steam-driven  pump  for  this  plant  would  consume  62  Ib.  of 
steam  per  hr.  It  would  consume  62  Ib.  of  steam  in  pumping  300  gal. 
boiler-feed  water. 

Now  the  steam  required  to  operate  an  equivalent  electrically-driven 
pump  will  be  determined:  A  test,  reported  by  the  Midvale  Machine 
Company,  indicates  that  360  gal.  of  boiler-feed  water  were  pumped,  in 
a  plant  similar  to  that  under  consideration,  with  an  energy  expenditure 
of  0.68  kw.-hr.  by  a  Johns  electrically-driven  pump.  The  following 
determination  will  be  based  on  the  data  of  this  test. 

In  the  50-kw.  plant  under  consideration,  it  is,  as  previously  stated, 
assumed  that  there  will  be  required  50  Ib.  of  steam  to  develop  1  kw.-hr. 
Hence,  to  develop  0.68  kw.-hr.  there  would  be  required:  0.68  X  50  =  34 
Ib.  of  steam.  Now  if  360  gal.  of  water  were  pumped  in  the  test  by  34  Ib. 
of  steam,  the  300  gal.  of  water,  required  per  hour  in  this  plant  would  be 
pumped  by  300  -r-  360  X  34  =  28.5  (approx.)  Ib.  of  steam.  . 

Now  the  saving  in  steam  due  to  the  use  of  the  .Johns  electrically-operated 
pump  will  be  the  difference  between  the  steam-pump  steam  consumption 
and  the  equivalent  electrically-driven  pump  steam  consumption.  Thus: 

Steam  pump  requires  per  hr.  (to  pump  300  gal.  feed 

water) 62.0  Ib.  of  steam 

Elec.  driv.  pump  requires  per  hr.  (to  pump  300  gal. 

feed  water) 28. 5  Ib.  of  steam 

Saving  in  steam  per  hr.  due  to  use  of  elec.  driven  pump     33 . 5  Ib.  of  steam 

Now  the  saving  in  dollars  due  to  the  use  of  the  motor-driven  pump  can 
be  determined:  The  boiler  evaporation  in  a  small  plant,  of  the  character 
of  that  under  consideration,  will  be  about  7  Ib.  of  water  for  each  pound  of 
coal.  Then,  the  coal  required  to  evaporate  33.5  Ib.  of  water  (the  amount 
which  is  saved,  each  hour,  by  the  use  of  the  electrically-driven  pump) 
will  be: 

Ib.  of  water  evaporated       33.5        .  0  „      -.        ,          L 

-   =  —   -  =  4.8  Ib.  of  coal  per  hr. 
rate  of  evaporation  7 

For  one  month  the  coal  saved,  based  on  this  50-kw.  load,  would  be 
30  days  X  24  hr.  X  4.8  Ib.  per  hr.  =  3,460  Ib.  per  month.  With  coal 
costing  $3.00  per  ton  in  the  bin,  the  saving  per  month  would  be:  (3,460  X 
3.00)  -T-  2,000  =  $5.19.  Or  the  saving  per  year  would  be,  approxi- 
mately: 12  X  $5.19  =  $62.25. 


SEC.  219] 


BOILER-FEEDING  APPARATUS 


189 


219.  Table  Showing  General  Advantages  And  Disadvan- 
tages For  Boiler  Feeding  Of  Power,  Electric,  and  Steam 
Pumps  When  Considered  Independently. 


A 

Mechanically  driven 

B 

Motor  driven 

C 

Steam  driven 

Advantages 


1.   Simplicity. 


2.  Low  cost  of  equipment. 


3.  Lowest  theoretical  cost 
of     operation     where     non- 
condensing    engine    is    used 
and  pump  is  belted  direct. 

4.  Cannot    run    away    in 
case  of  accident. 


1.  May  be  easy  of  con- 
trol.    (Automatic     control 

e.) 

2.  Fairly  simple. 


3.   More     efficient     than 
steam — low  cost  per  year. 


4.  Location  where  desir- 
able. 

5.  Can    be    operated  in- 
dependently  of  boiler  and 
engine     unit     under     some 
circumstances. 

6.  In    case    of    accident, 
has   speed    limit,    i.e.,    will 
not  "run  away." 

7.  Accurate  heat  balance 
possible  with  proper  equip- 
ment.    (Sec.  212.) 


1.  Ease  of  control. 


2.  Supply  of  exhaust 
steam  for  feed-water  heat- 
ing in  condensing  plant. 

3.  Low  cost  of  equipment. 


4.  Maximum  reliability. 


Disadvantages 


1 .  Poor  regulation  of 
water   supply — extra    water 
rnay    be    pumped    and    re- 
turned if  by-pass  is  used. 

2.  May    be    unhandy    in 
location. 


3.  Works  only  when  main 
engine  is  running. 


1.  Cannot  be  used  with- 
out     generator.        (Unless 
Public- Service-Company 
power  is  available.  *) 

2.  In    condensing    plant 
increase     loss     of     exhaust 
steam   heat    in    condensing 
water  unless  pump  is  driven 
from      auxiliary      non-con- 
densing engine  unit  exhaust 
from  which  is  used  for  feed- 
water  heating. 


1.  Great  loss  due  to  in- 
efficiency    unless      exhaust 
steam  can  be  utilized. 

2.  Possible    self    destruc- 
tion   in    case    of    feed-line 
breaking. 


*  Where  Public-Service-Company  power  is  available  this  feature  may  be  important. 


190 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


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SEC.  220] 


BOILER-FEEDING  APPARATUS 


191 


fen  pumps. 

fen  pumps. 

fen  pumps. 

fen  pumps. 

fen  pumps. 

fen  pumps. 

team-driven  pumps. 

Ise  the  one  which  will  provide  better  heat 

Use  A  normally  and  C  for  stand-by. 

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>team-driven  pumps. 

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192  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

221.  A    Turbine-    Or    Motor -Driven    Centrifugal    Pump 
(see    Div.   4)    affords,   ordinarily,   the  best  unit  for  regular 
operation  for  pumping  boiler  feed  water  for  plants  of  capaci- 
ties exceeding  about  500  h.p.     A  500-h.p.  plant  is  equivalent 
to  a  feed-water  requirement  of  about  50  gal.  per  min.  or  3,000 
gal.  per  hr.     (However,  in  every  case  there  should  be  a  steam 
direct-acting  stand-by  pump.)     The  centrifugal  pump  is  the 
best  for  this  service  because  it  will,  in  the  long  run,  prove  the 
most  economical.     It  has  the  advantage  that  the  discharge 
from  the  pump  to  the  boiler  may  be  throttled  down  or  opened 
as  desired   without  the   considerable  loss  of  energy  which 
results  from  by-passing. 

NOTE. — The  pressure  developed  by  a  centrifugal  pump  which  is  oper- 
ated at  normal  speed  can  never  exceed  a  certain  maximum.  Further- 
more, if  the  feed  line  from  the  pump  to  the  boiler  should  break,  thus 
reducing  the  head  against  which  the  pump  is  forcing  water  to  practically 
zero,  the  centrifugal  pump  will  not  "run  away,"  but  it  will  continue  to 
operate  at  practically  constant  speed.  Its  power  consumption  will  be  very 
low  when  it  is  pumping  against  zero  head.  Again,  if  the  valve  in  the 
discharge  pipe  in  a  centrifugal  pump  is  closed  the  pump  may  continue 
to  turn  at  its  normal  speed  (Sec.  171)  without  developing  an  excessive 
water  pressure.  In  such  a  case  the  water  is  merely  churned  around 
within  the  casing. 

222.  The   Efficiency   Of   The  Centrifugal  Pump  remains, 
with  slight  repair,  nearly  constant  throughout  its  life  because 
there  is  practically  nothing  about  it  except  two  simple  bearings 
to  wear  out.     Obviously,  where  gritty  water  is  being  pumped 
through  there  will  also  be  wear  on  the  impellers  or  blades,  but 
gritty  water  is  not  used  for  boiler  feed.     On  the  other    hand, 
the  efficiency  of  any  plunger  type  or  piston  pump  may  decrease 
decidedly  as  the  pump  becomes  older,  due  to  leaky  valves, 
pistons,  and  worn  rods.     This  is  true  particularly  of  the  single 
or  duplex  steam  pump.     The  water  rate  of  such  a  steam  pump 
after  a  year  or  so  of  service  and  insufficient  maintenance  may 
be  twice  its  initial  water  rate. 

223.  The  Centrifugal  Pump  Has  No  Valves    Which    He- 
quire  Re-Grinding. — Unfortunately,  the  valves  of  any  plunger 
or  piston  pump  do  not  usually  receive  the  attention  which 
they   should   have.     With   the   steam   pump,    if   the   valves 
become   leaky   the   operator   may    merely    "give   her   more 


SEC.  224]  BOILER-FEEDING  APPARATUS  193 

steam."  Thus,  the  required  water  may  be  pumped,  but 
uneconomically.  Such  losses  are  difficult  to  locate  because 
the  steam  requirements  of  the  boiler  feed  pumps  are  such  a 
small  proportion  of  the  total  steam  requirements  of  the  plant. 

NOTE. — THE  WATER  RATE  OF  A  TURBINE  FOR  DRIVING  A  SMALL 
CENTRIFUGAL  PUMP  will  be  from  38  to  43  Ib.  of  steam  per  brake  h.  p.  hr. 
This  consumption  does  not  increase  materially  as  the  age  of  the  turbine 
increases. 

NOTE. — THE  MECHANICAL  EFFICIENCY  OF  A  CENTRIFUGAL  PUMP 
(the  capacities  range  from  50  gal.  per  min.  and  up)  will  be  from  50  to 
60  per  cent.  For  the  larger  centrifugal  pumps  operating  under  favorable 
conditions  efficiencies  as  high  as  81  per  cent,  have  been  obtained. 

224.  One  Disadvantage  Of  Centrifugal  Pumps  for  boiler- 
feeding  is  the  fact  that  if  the  feed  water  is  very  near  its  boiling 
point,  the  action  of  the  pump  may  vaporize  it  entirely  within 
the  pump  casing  (see  Sec.  157).     If  this  happens,  the  pump 
will  not  work  as  it  depends  on  the  action  of  the  runner  on  a 
liquid.     On  the  other  hand,  a  plunger  pump  will  handle  water 
at  any  temperature  as  long  as  there  is  pressure  enough  to 
deliver  the  water  to  the  pump  cylinders.     Moreover  a  cen- 
trifugal pump   cannot  be  run  at  all  if  in  poor  condition,  on 
account  of  its  high  speed.     If  there  is  any  damage  to  shaft 
or  runner,  the  pump  must  usually  be  shut-down  and  com- 
pletely overhauled.     It  is  for  these  reasons  that  a  centrifugal 
pump  is  not  recommended  in  this  Div.  for  a  stand-by  pump. 

225.  Power  Boiler -Feed -Pump  Sizes  For  Various   Boiler 
Horse  Powers  as  taken  from  The  Goulds  Mfg.  Go's,  catalogue 
are  given  in  the  two  tables  which  follow.     The  tabulated 
values  indicate  the  water  supply  required  by  the  boiler  based 
on  the  A.  S.  M.  E.  standard  rating  (Sec.  229)  of  34>^  Ib.  of 
feed  water  per  boiler  horse  power  hour  from  and  at  212  deg. 
fahr.     A  surplus  of  25  to  50  per  cent,   pump  capacity   is 
recommended.     See  Sec.  228  for  methods  of  computing  boiler 
feed-water  requirements. 

226.  Table    Showing   Boiler-Feed   Capacities   Of   Single- 
Acting  Triplex  Power  Pumps.     Goulds  Mfg.  Co.  (See  preced- 
ing Sec.).     The  capacity  of  a  double-acting  simplex  pump  is 
approximately  0.66  times  of  that  tabulated  for  the  same  speed 
and  cylinder  dimensions.     The  capacity  of  a  double-acting 

13 


194 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


duplex  pump  is  1.33  times  that  tabulated  for  the  same  speed 
and  cylinder  dimensions. 


Rated  capacity  of 
boilers, 
horse  power 

Feed  water  at 
212°F., 
Gallons  per 
minute 

Size  of  pumps, 
inches 

Revolutions 
per  minute 

30 

2.15 

IH  X  2M 

30 

50 

3.59 

2      X  3 

31 

100 

7.17 

2^  X4 

30 

150 

10.75 

3      X  4 

31 

200 

14.34 

3K  X4 

30 

400 

28.7 

4      X  6 

31 

800 

57.4 

5       X8 

30 

1200 

86. 

6      X  8 

31 

1600 

115. 

7      X  8 

30 

2000 

143.4 

7      X  8 

30 

2750 

196. 

8      X  10 

31 

4000 

286. 

9      X  12 

30 

5000 

358. 

10      X  12 

30 

227.  Table    Showing   Boiler-Feed    Capacities    Of    Multi- 
stage Centrifugal  Pumps.     Goulds  Mfg.  Co.  (See  Sec.  225). 


Horse 
power  of 
boilers, 
rated 
capacity 

Feed 

water  at 
212°F., 
gallons 
per 
minute 

Size  of 
pipe 
dis- 
charge, 
pipe 
inches 

Revo- 
lutions 
per 
minute 

Horse 
power  of 
boilers, 
rated 
capacity 

Feed 

water  at 
212°F., 
gallons 
per 
minute 

Size  of 
pipe 
dis- 
charge, 
pipe 
inches 

Revo- 
lutions 
per 
minute 

700 

50.20 

2 

3500 

2800 

200.5 

4 

2500 

850 

60.92 

2 

3500 

3150 

225.9 

4 

2500 

1000 

71.60 

2 

3500 

3500 

250.8 

4 

2500 

1200 

86.00 

2 

3500 

3850 

275.8 

4 

2500 

1500 

107.50 

2 

3500 

4200 

301.0 

4 

2500 

1750 

125.  30 

2 

3500 

4900 

351.0 

5 

2200 

2000 

143.33 

3 

3100 

5600 

402.0 

5 

2200 

2100 

150.50 

3 

3100 

6300 

452.0 

5 

2200 

2450 

175.  50 

3 

3100 

7000 

502.0 

5 

2200 

228.  There  Are  Two  Methods  Of  Estimating  Feed-Water 
Requirements  Of  A  Power  Plant. — One  is  based  on  the  rating 
of  the  boilers  in  the  plant.  The  other  is  based  on  the  actual 
steam  consumption  of  the  engines  and  auxiliaries  or  devices 


SEC.  229]  BOILER-FEEDING  APPARATUS  195 

for  which  the  steam  is  generated.  Which  method  should  be 
used  in  any  case  will  be  determined  by  conditions.  Probably 
the  second  method,  that  based  on  the  actual  steam  consump- 
tions, is  the  more  accurate.  But,  in  a  plant  in  which  steam 
is  used  only  for  power  generation,  if  the  boiler  capacities  are 
proportioned  rationally  in  relation  to  the  units  which  they 
supply,  both  methods  should  give  approximately  the  same 
results.  In  ascertaining  the  feed-water  requirements  for  a 
power  plant  it  may  be  wise  to  make  an  estimate  by  each  of  the 
methods,  compare  the  results  as  a  check,  and  then  take  for  a 
working  basis  the  one  which  is  the  larger.  Where  a  boiler 
plant  generates  steam  for  heating  only,  that  is,  where  there 
are  no  steam-consuming  units,  such  as  pumps  and  engines,  it 
is  obvious  that  then  only  the  first  method,  that  based  on  the 
boiler  rating,  is  applicable. 

229.  In  Determining  Feed-Water  Requirements  On  The 
Basis  Of  The  Boiler  Rating  the  accepted  water-rate  equiva- 
lent of  a  boiler  horsepower  (boiler  h.p.)  is  utilized.  The 
equivalent  is  this:  It  was  recommended  by  the  American 
Society  of  Mechanical  Engineers  in  1899  that  the  evaporation 
of  34.5  Ib.  of  water  per  hr.  at  212  deg.  be  taken  as  the  equiva- 
lent of  1  boiler  h.p.  This  equivalent  is  now  universally 
accepted  as  standard  in  the  United  States.  Hence,  the  process 
of  determining  the  water  required  to  feed  a  boiler  is  this: 
(1)  Ascertain  the  total  h.p.  rating  of  the  boiler  or  boilers  in 
question.  (2)  Multiply  this  total  h.p.  rating  by  34.5  which 
will  give  the  number  of  pounds  of  water  required  per  hour 
when  the  boiler  is  operated  at  rated  capacity.  (3)  Now  due 
to  the  fact  that  the  pump  must  occasionally  raise  the  water 
level  and  pump  more  than  34.5  Ib.  per  hr.,  the  value  obtained 
in  this  manner  should  be  increased  by  30  to  50  per  cent. 

In  fact,  a  boiler  feed  pump  is  usually  selected  on  the  basis 
that  it  will  deliver  45  to  50  Ib.  of  water  per  hr.  for  each  rated 
boiler  h.p.  The  operations  above  applied  may  be  expressed 
in  a  formula,  thus: 

(71)  Lb.  of  water  per  hr.  =  Wwh  X  PBhP 

Wherein :  Wwh  =  the  Ib.  of  water  per  boiler  h.p.  hr.  upon  which 
the  estimate  is  based.  This  value  may  vary  from  45  to  50. 


196  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

The  value  of  45  Ib.  per  hr.  is  conservative  and  may  ordinarily 
be  assumed.  PBhP  =  the  total  rated  boiler  h.p.  of  the  boiler 
or  boilers  which  are  to  be  fed.  Now  since  there  are  8.34  Ib. 
of  water  in  a  gallon,  it  follows  that: 

(72)  Gal.  required  per  hr.  =  W»*  *  P*** 

o.o4 

Now  if  Wwh  be  taken  as  45,  then 

(73)  Gal.  required  per  hr.  =  45  X  PBhP  •*-  8.34  =  5.4  PBhp. 

which  is  the  accepted  working  formula.  Where  Wwh  is  taken 
as  50  Ib.: 

(74)  Gal.  required  per  hr.  =  6  X  PBhP. 

EXAMPLE.— A  boiler  has  500  rated  h.p.  What  should  be  the  capacity 
of  the  feed-water  pump  to  supply  it?  SOLUTION. — Base  the  estimate  on 
45  Ib.  of  water  per  rated  h.p.  hr.  Then  substitute  in  the  above  formula, 
thus:  Gal.  required  per  hr.  =  5AXPBhP  =  5.4 X  500=  2,700. 

Hence,  a  pump  capable  of  delivering  at  least  2,700  gal.  of  water  per 
hour  should  be  installed. 

230.  In  Determining  The  Feed-Water  Requirements  Of  A 
Power  Plant  On  The  Basis  Of  Its  Steam  Consumption,  the 

process  is  this:  (1)  Ascertain,  either  from  manufacturers' 
guarantees,  or  by  using  a  table  of  water  rates,  the  pounds  per 
hour  of  steam  required  for  the  engine  or  principal  units.  (2) 
Similarly  determine  the  pounds  of  steam  required  per  hour  by 
the  auxiliaries.  Then  disregarding  radiation,  leakage,  steam 
required  by  the  whistle,  and  other  losses: 

(75)  Total  weight  of  water  required  per  hr.  =  (1)  +  (2) 

To  allow  for  the  radiation,  leakage,  whistle  loss,  and  to 
provide  some  capacity  for  forcing  and  for  recovering  the 
water  level  in  case  it  is  lost,  the  value  obtained  by  the  equation 
just  above  should  be  increased  by  25  per  cent. 

EXAMPLE. — What  will  be  the  probable  feed-water  requirement  of  a 
plant  which  operates  a  50-h.p.  high-speed  condensing  engine  and  10  h.p. 
of  non-condensing  auxiliaries?  SOLUTION. — First  determine  the  water 
consumption  of  engine  and  auxiliaries.  From  a  table  of  water  rates  it  is 
found  that  a  50-h.p.  high-speed  condensing  engine  will  have  a  water  rate 
of  about  22  Ib.  per  h.p.  hr.  Hence,  its  total  steam  consumption  will  be: 
(50  X  22)  =  1,100  Ib.  per  hr.  The  water  rates  of  the  small  auxiliaries 


SEC.  231]  BOILER-FEEDING  APPARATUS  197 

will  probably  be  200  Ib.  per  h.p.  hr.  Hence,  the  total  auxiliary  steam 
consumption  will  be:  (10  X  200)  =  2,000  Ib.  of  steam  per  hr.  Steam  con- 
sumption of  engine  and  auxiliaries,  then,  is:  (1,100  +  2,000)  =  3,100. 
Multiplying  this  by  1.25  to  allow  for  losses  and  forcing,  thus:  (3,100 
X  1.25)  =  3,875  Ib  .  This  is  the  total  weight  of  water  required  per 
hour.  To  reduce  this  to  gallons,  divide  by  8.3,  thus  (3,875  -5-  8.3)  = 
467  gal.  per  hr. 

231.  Increased  Feed-Pump  Capacity  Is  Necessary  If  The 
Modern  Large-Plant  Practice  Of  Forcing  The  Boilers  Is 
Followed. — In  large  power  plants  where  automatic  stokers 
can  be  used,  particularly  if  the  plant  is  situated  in  a  city, 
boilers  are  " forced"  so  that  their  output  much  exceeds  the 
nominal  evaporation  of  34.5  Ib.  per  rated  boiler  h.p.  per  hr. 
by  possibly  150  to  200  per  cent.  A  forced  boiler  is  not  as 
efficient  as  one  which  is  being  worked  conservatively.  This 
is  because  that,  when  a  boiler  is  forced,  a  larger  proportion  of 
the  heat  of  the  coal  is  wasted  in  the  flue  gases  than  if  the  boiler 
is  not  forced.  That  is,  the  flue  gas  temperatures  in  the  smoke 
stack  will  be  higher  in  the  case  of  the  forced  boiler.  This  is 
equivalent  to  a  loss.  But  in  spite  of  the  fact  that  the  boiler 
efficiency  is  decreased  when  the  boiler  is  forced,  it  usually 
works  out  in  the  larger  power  plants  that  it  is  more  economical 
to  force  the  boilers  than  it  would  be  to  pay  the  additional 
fixed  charges  on  boiler  investment,  maintenance  and  real 
estate  that  would  be  involved  if  sufficient  boiler  capacity  were 
installed  to  insure  operation  on  the  basis  of  an  evaporation  of 
34.5  Ib.  of  water  per  rated  boiler  h.p.  per  hr. 

NOTE. — THE  LIFE  OF  A  BOILER  WHICH  Is  BEING  FORCED  may  not, 
unless  the  forcing  is  extremely  excessive,  be  less  than  that  of  a  boiler 
which  is  not  forced.  It  is  essential,  however,  that  a  forced  boiler  be 
provided  with  purified  feed  water,  otherwise  scaling  and  blistering 
difficulties  are  bound  to  occur. 

NOTE. — IN  PRACTICE,  BOILERS  IN  LARGE  PLANTS  ARE  Now  OFTEN 
FORCED  To  150  To  200  PER  CENT,  of  the  A.  S.  M.  E.  rating  for  normal 
operation  and  at  peak  load  they  may  be  forced  to  300  per  cent,  rating. 
That  is,  for  each  rated  boiler  h.p.  (on  the  A.  S.  M.  E.  basis  of  an  evapora- 
tion of  34.5  Ib.  of  water  per  hr.)  a  boiler  which  is  being  forced  to  150  per 
cent,  of  its  rating  will  then  evaporate:  34.5  X  1.5  =  51.75  Ib.  of  water  per 
hour.  Similarly,  if  a  boiler  is  being  forced  to  200  per  cent,  rating  it 
will  evaporate  69  Ib.  of  water  per  rated  boiler  h.p.  per  hour.  At  peak 
load  periods  when  forced  to  300  per  cent,  rating,  the  evaporation  may  be 


198 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


Pipe  Connect/or? 
To  Feed  Line 


103.5  Ib.  of  water  per  rated  boiler  h.p.  per  hour.  So  it  is  evident  that 
the  rule  given  in  one  of  the  opening  paragraphs  of  this  section  for  com- 
puting feed-pump  capacity  on  the  basis  of  rated  boiler  h.p.  may  have 
to  be  modified  materially  if  the  pump  is  to  be  used  in  the  plant  where 
the  boilers  are  forced.  In  such  plants  the  safer  procedure  is  to  determine 
the  actual  steam  consumptions  per  hour  of  the  prime  movers  and  of  all 
of  the  auxiliaries  and  base  the  feed-pump  rating  on  this  total  steam  con- 
sumption. Furthermore,  in  estimating  boiler-feed-pump  capacities, 
ample  allowance  should  be  made  for  future  additions  to  the  boiler 
equipment,  if  any  are  contemplated. 

232.  Pump  Governors  On  Direct-Acting  Steam  Pumps  In 
Boiler-Feed  Service  (Fig.  210)  operate  in  conjunction  with 

feed-water  regulators.  (See  the 
author's  STEAM  BOILERS.)  The 
function  of  the  pump  governor  is 
to  maintain  a  constant  pressure 
in  the  feed  line.  It  does  this  by 
moderating  the  speed  of  the 
pump,  or  shutting  it  down  alto- 
gether, when  the  feed-water  reg- 
ulator diminishes  the  openings 
through  the  feed  valves  or  closes 
them  entirely.  If  no  governor 
were  used  to  regulate  the  speed 
of  the  pump  in  response  to  the 
feed- water  regulator's  adjust- 
ment of  the  feed-valves,  the 
pump  might  build  up  a  pressure 
in  the  feed-line  powerful  enough 
either  to  force  an  excess  quan- 
tity of  water  into  the  boilers 

through  the  partially  closed  feed- 
Em.  210.— Sectional  Elevation  Of  ,          .    . 
Fisher    Pump    Governor  For  Boiler-     ValVCS  Or  to  burst  the  piping.       A 

Feed  Pumps.  properly  working  pump  governor 

controls  the  movement  of  the  pump  piston  or  pistons  so  as  to 
constantly  maintain  a  pressure  in  the  feed  line  just  enough 
greater  than  the  boiler  pressure  to  insure  a  positive  flow  of 
the  water  into  the  boilers  against  the  steam  pressure. 

EXPLANATION. — The  Fulton  pump  governor  (Fig.  211)  is  connected 
into  the  live  steam  supply  at  B  and  to  the  steam  end  of  a  direct-acting 


SEC.  232] 


BOILER-FEEDING  APPARATUS 


199 


steam  pump  at  A .  The  small  pipe  E  is  connected  to  the  feed  line  so  that 
the  under  side  of  diaphragm  D  is  subjected  to  feed  water  pressure.  The 
upper  side  of  D  is  subjected  to  live  steam  pressure  (or  approximately 
boiler-pressure)  through  the  passage  C.  The  vertical  stem  of  valve  V  is 
acted  upon  by  weight  W  at  one  end  and 
diaphragm  D  at  the  other  When  the 
feed  line  pressure  is  less  than  boiler  pres- 
sure, both  the  weight  and  the  diaphragm 
tend  to  open  the  balanced  valve  V.  Then 
steam  flows  freely  through  the  governor 
from  B  to  A  and  operates  the  pump  at 
full  speed.  The  feed  line  pressure  is  built 
up  by  the  pump  untiL  the  weight  W  is 
lifted  and  valve  V  closed  by  the  pressure 
on  the  under  side  of  the  diaphragm.  The 
weight  may  be  adjusted  to  open  the  valve 
at  any  desired  pressure. 

EXAMPLE. — Suppose  the  boiler  pressure 
is  100  Ib.  per  sq.  in.  and  a  pressure  of 
110  Ib.  per  sq.  in.  in  the  feed  line  is  satis- 
factory for  delivering  water  against  the 
pressure  of  the  boiler.  Therefore  when 
the  pump  is  working  at  the  proper  rate, 
there  will  be  110  Ib.  per  sq.  in.  on  the 
under  side  of  the  diaphragm  (D  Fig.  211) 
and  100  Ib.  per  sq.  in.  on  the  upper  side. 
The  weight  W  is  set  so  as  to  overcome  the 
force  of  this  difference  in  pressure  of  (110  —  100)  or  10  Ib.  per  sq.  in. 
Therefore  when  the  difference  in  pressure  is  a  little  less  than  10  Ib.  per 
sq.  in.,  the  weight  opens  the  valve  V.  When  the  difference  in  pressure 
is  a  little  more  than  10  Ib.  per  sq.  in.,  the  diaphragm  closes  the  valve. 
In  this  way  a  pressure  difference  just  sufficient  to  feed  the  boiler  is 
maintained. 


FIG.  211. — Sectional  Elevation 
Of  The  Fulton  Governor  For 
Boiler-Feed  Pumps. 


Fia.  212. — Section  Of  Horizontal  Piston  Type  Pump  Governor. 

NOTE. — IN  THE  HORIZONTAL  TYPE  OF  PUMP  GOVERNOR  (Fig.  212), 
the  piston  P  takes  the  place  of  the  diaphragm  and  a  spring  S  takes  the 
place  of  the  weight  as  described  above.  The  piston,  however,  is  acted  on 
by  the  feed  line  pressure  only  and  so  communicates  this  full  pressure  to  the 
spring.  The  tension  on  the  spring  is  adjusted  by  means  of  two  thumb- 
screws T. 


200 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


233.  The   Fisher  Pump  Governor   (Figs.  210  and  213)  is 
similar  to  the  Fulton  governor  in  operation.     One  advantage 


Pax  Convzuinq  Steam  Pressure 


%>  Under  5/ofe  Of  Diaphragm- ~ 


>  \\\\\\\\\\\\V 
'  Auction  Ploe  From  Teed-  Wafer  Hecrfer 


FIG.  213.  —  Direct-Acting  Boiler-Feed  Pump  Equipped  With  Fisher  Governor. 


Steam  Main--'*' 

Steam  Line  7i 
f  Diaphragm- - 


from  Md/'n  Steam  Line  To  Under  Side 
Of  D 


.-Main  Feed  Line  To  Boilers 

\    From,  Main  Feed  Line  To  Upper 
\         Side  Of  Diaphragm-^ 
*  \ 


Steam 
Supply 

To 
Pump-, 


Governor  Spriny  S- 
Governor  Valve- ^~... 


Suction  Pipe  From 
Feed  Worter  Heater- 


FIG.  214.— Direct-Acting  Boiler  Feed  Pump  Equipped  With  Kieley  Governor. 

of  the  Fulton  design  shown  is  that  it  uses  only  one  stuffing 
box  where  the  Fisher  design  requires  two.     One  advantage 


SEC.  234]  BOILER-FEEDING  APPARATUS  201 

of  the  Fisher  design  is  that  the  steam  pressure  can  be  shut  off 
from  the  diaphragm  chamber  for  inspection  and  repair.  The 
Kieley  governor  (Fig.  214)  uses  a  spring  in  connection  with  a 
diaphragm  in  chamber  D. 

NOTE. — Pump  governors  for  maintaining  constant-pressure  are  some- 
times used  on  turbine-driven  centrifugal  boiler-feed  pumps  (Fig.  204). 
Their  function  is  merely  to  save  steam,  as  there  is  little  danger  from  the 
over-pressure  of  a  centrifugal  pump.  A  centrifugal  over-speed  governor 
may  be  provided  on  the  same  unit  as  a  constant-pressure  governor. 

234.  A  Water-Relief  Valve  Must,  Where  A  Feed-Water 
Regulator  Is  Used,  Be  Installed  On  A  Constant-Speed  Crank- 
Action  Feed  Pump. — The   water-relief  valve    (Fig.    199)   is 
merely  a  special-type  safety  valve.     Crank-action  feed  pumps 
usually  run  at  fairly-constant  speed  (Sec.  217)  and  are  always 
pumping  about  the  same  amount  of  water.     Hence,  if  the 
feed-water  regulator  partially  or  wholly  closes  the  feed-water 
line  to  the  boilers,  stalling  of  the  pump  or  damage  to  the  pump 
and  its  accessories  are  liable  to  result  unless  a  water-relief 
valve  is  provided  to  automatically  by-pass  the  surplus  water. 

NOTE. — A  WATER-RELIEF  VALVE  SHOULD  BE  PROVIDED  IN  THE  BY- 
PASS ON  EVERY  RECIPROCATING  POWER  FEED  PUMP  as  a  safety  measure, 
whether  or  not  a  feed  water  regulator  is  employed.  This  is  to  prevent 
damage  if  the  feed-water  line  to  the  boilers  is  closed  accidentally. 

235.  The  Most  Common  Troubles  Of  Pump    Governors 
And  Their  Causes  And  Remedies  are  as  follows : 

1.  BLOWS  STEAM  AROUND  VALVE  STEM.  Should  be  entirely  re- 
packed with  fine  packing  and  lubricated  with  cylinder  oil  and  graphite. 
Screwing  up  the  packing  gland  to  stop  steam  leaks  is  likely  to  make  too 
much  friction  before  the  gland  is  tight. 

2.  Too   SLUGGISH — gives  too  much  variation  in  feed-line  pressure. 
Friction  in  the  movement  usually  gives  this  effect.     Sometimes  the 
spring  used  is  too  stiff  for  the  pressure  in  governors  of  the  spring  type. 
See  if  the  valve  stem  slides  freely.     If  it  does  not,  the  friction  must  be 
located  and  then  remedied  by  polishing  and  lubrication.     Sometimes  a 
stuffing  box  is  too  tight  or  the  packing  old  and  stiff.     A  weaker  spring 
gives  less  variation  of  feed-line  pressure. 

3.  GIVES  CONSTANT  PRESSURE  Too  Low  OR  Too   HIGH.      Adjust 
weight    or  spring  thumb  screws.     Increase  spring  tension  or    weight 
leverage  for  more  pressure. 

4.  DOES  NOT  SHUT  OFF — gives  excessive  feed-line  pressures  as  shown 
by  gage  or  by  creeping  or  other  signs  of  overpressure  in  pump. 


202 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.6 


(a)  Friction  in  stem  or  piston.     Remedy  as  explained  before. 

(6)  Damaged  diaphragm.  Remove  this  member.  A  high-grade  rub- 
ber packing  re-inforced  with  several  layers  of  fabric  may  be  used  for 
diaphragms  but  plain  rubber  is  not  suited  for  the  purpose.  An  extra 
diaphragm  should  be  ordered  from  the  manufacturer  and  kept  on  hand. 

(c)  Valve  does  not  seat.  Examine  seat  for  scores  and  corrosion.  If 
it  seems  to  be  in  fair  condition  grind  with  grinding  compound  and  see  if 
a  clean  face  can  be  obtained.  If  scored  too  deeply  to  be  ground  clean, 
the  valve  must  be  re-finished  on  a  lathe.  In  re-finishing,  the  angle  of 
the  face  and  span  between  the  two  faces  must  be  accurately  retained. 
After  finishing,  the  valve  should  be  " ground  in"  and  all  grinding  com- 
pound removed  before  re-assembling. 

236.  Automatic  Apparatus  For  The  Feeding  Of  Boilers 
With  Hot-Water  Returns  from  heating  systems  are  of  two 


Counter-weight  For  Closing  Stec/m.  Votive 


Vent 
To  Roof-, 


Suction  Connection 
ToStanof  Pipe..... 


•Outlet  Jo, 
Stand  Pipe 


FIG.  215. — Duplex  Steam  Pump  And  Receiver  Arranged  For  Automatic  Return,  To 
Boiler,  Of  Condensate  From  Heating  Apparatus. 


principal  types:  (1)  The  combined  pump  and  receiver  (Fig.  215). 
(2)  The  return  trap  (Fig.  216).  With  both  classes  of  apparatus 
the  hot  water  or  condensate  which  returns  from  the  radiators 
and  heating  coils  is  collected  in  a  receiving  tank.  By  the 
first  method,  however,  a  direct-acting  steam  pump  is  auto- 
matically operated  to  discharge  the  water  from  the  receiving- 
tank  into  the  boiler.  By  the  second  method  the  water  is 
dumped  directly  from  the  receiving-tank  into  the  boiler. 


SEC.  236] 


BOILER-FEEDING  APPARATUS 


203 


EXPLANATION. — WITH  THE  COMBINED  PUMP  AND  RECEIVER  (Fig.  215) 
the  condensate  from  the  heating  apparatus  enters  the  receiver  through  the 
inlet  nozzle.  When  the  body  of  water  accumulates  until  its  surface 
stands  at  about  half  the  height  of  the  receiver  it  buoys  up  the  bucket- 
float  B.  Steam  is  thereby  admitted  to  the  pump  through  the  valve  V, 
the  stem  of  which  is  connected  to  the  float-lever  at  F.  As  the  water- 
level  in  the  receiver  is  lowered  by  the  action  of  the  pump  the  opening 


P/pe  Connection 
Upper  And  Low&r 


Lowzr  Kece/Ver-' 
Wer  Trap 

Fio.  216. — Bundy  Traps  Arranged  For  Return  To  Boiler  Of  Condensate  From  Heating 

Apparatus. 

through  the  steam  valve  V  is  gradually  diminished,  due  to  the  depression 
of  the  float.  The  speed  of  the  pump  is  thus  regulated  according  to  the 
quantity  of  water  flowing  into  the  receiver.  The  water  which  is  required 
to  make  up  for  loss  of  steam  or  condensate  from  the  system,  due  to  leak- 
age or  other  cause,  is  admitted  at  M. 

WITH  THE  RETURN-TRAP  METHOD  (Fig.  216)  the  condensate  from  the 
heating  system  collects  in  the  lower  receiver,  Rlt  and  flows  thence  into 
the  bowl,  Bi,  of  the  lower  trap  T\.  When  sufficient  water  has  accumu- 
lated in  the  bowl  BI  to  cause  it  to  tilt  (Sees.  487  and  488)  steam  at  boiler 


204 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  6 


pressure  enters  through  the  pipe  PI  and  forces  the  water  into  the  upper 
receiver,  R2,  whence  it  flows  into  the  bowl  B2  of  the  upper  trap  TV 
This  trap  is  located  3  ft.  or  more  above  the  normal  water-level  in  the 
boiler.  When  the  bowl  B 2  tilts  under  the  weight  of  the  accumulated 
water,  steam  at  boiler  pressure  enters  through  the  pipe  PI.  The  pres- 
sure in  the  trap  and  in  the  boiler  is  thus  equalized.  Due  to  its  static 
head  of  3  ft.  or  more,  the  water  in  the  bowl  B2  flows,  by  gravity,  into  the 
boiler  through  the  feed-pipe  F.  When  empty,  the  bowl  tilts  back  to  its 
filling  position. 

237.  The  Duplex  Boiler-Feeder  (Figs.  217  and  218)  operates 
similarly  to  a  return  trap  system  but  has  larger  capacity. 


Equalizing  P/'pes- 


Wcrf-er  Conmc  tiom : 
Of?  Reverse  End—' 

Exhaust  Sfectm 
FIG.  217. — Farnsworth  Duplex  Boiler  Feeder. 

This  feeder  is  recommended  by  its  manufacturers  for  boiler- 
feeding  in  non-condensing  plants  where  water  from  the  mains 

Ghonmbar-%      -Chamber-fr 

3=mi-u 


FIG.  218. — Showing  Installation  Of  Duplex  Boiler  Feeder  In  Connection  With  Closed 
Heater  In  Non-Condensing  Plant. 

is  fed  to  the  boiler  through  some  sort  of  feed-water  heater. 
It  depends  for  its  operation  on  a  water  supply  under  sufficient 
pressure  to  flow  to  the  top  of  the  boiler. 


SEC.  238]  BOILER-FEEDING  APPARATUS  205 

EXPLANATION. — The  feeder  shown  in  Fig.  217  is  located  above  the 
boiler.  The  tank  consists  of  two  equal  compartments  A  and  B  separated 
by  a  central  wall.  It  is  pivoted  below  its  center  of  gravity  so  that  it 
may  oscillate  a  few  degrees  in  either  direction.  A  system  of  valves  is 
arranged  in  the  pivot  so  that  whichever  compartment  is  down  is  allowed 
to  drain  into  the  boiler.  Boiler  pressure  is  admitted  at  the  top  of  the 
compartment  to  make  this  possible.  Meanwhile,  the  raised  compart- 
ment fills  with  water  from  the  feed  line.  Whenever  the  weight  of  water 
in  the  upper  compartment  is  sufficiently  greater  than  that  in  the  lower, 
the  tank  tilts  and  the  process  in  the  two  compartments  is  reversed. 

238.  The  Relative  Merits  Of  Pumps  And  Steam  Traps  For 
Boiler  Feeding  are  as  follows:  (Power  Plant  Engineering, 
Dec.  1,  1920).  Where  direct-return  steam  traps  can  be  used  to 
feed  a  boiler  or  boilers,  they  usually  provide  a  more  economical 
method  than  do  steam  pumps;  this  all  depends,  however,  on 
the  conditions  in  the  plant. 

EXPLANATION. — Where  returns  from  a  heating  system  are  fed  into  a 
boiler,  unless  the  boiler  is  low  enough  so  that  the  returns  can  feed  by 
gravity  to  a  trap  located  4  or  5  ft.  above  water  level  in  the  boiler,  it  is 
necessary  to  use  two  traps;  one  to  force  the  water  up  to  the  trap  above 
the  boiler  by  means  of  boiler  pressure  steam;  the  other  a  direct  return 
trap  to  dump  water  into  the  boiler.  In  such  a  case,  the  cost  of  a  trap 
installation  is,  of  course,  higher  than  that  for  a  pump,  which  can  force 
the  water  directly  into  the  boiler  without  rehandling. 

Where  the  feed  water  of  the  boiler  is  not  made  up  entirely  of  condensed 
steam,  and  the  load  is  variable  so  that  the  amount  of  feed  must  be 
varied,  the  speed  of  the  feed  pump  can  be  controlled  more  easily  than  a 
trap.  The  trap  simply  dumps  into  the  boiler  whatever  water  comes  to  it, 
and,  of  course,  the  rate  of  flow  into  the  trap  could  be  regulated  by  the 
valve  in  the  supply  line.  Where  cold  water  is  used  for  feed  and  has  to 
be  heated,  it  is  difficult  to  arrange  the  system  so  as  to  feed  through  a  trap, 
as  either  the  feed-water  heater  must  be  located  above  the  trap  or  a  lifting 
trap  be  employed  to  take  water  up  to  the  direct-return  trap.  The  chief 
argument  for  the  pump  is  convenience  and  flexibility,  and  adaptability 
to  all  conditions. 

QUESTIONS  ON  DIVISION  6 

1.  Name  the  three  principal  kinds  of  devices  used  in  boiler-feeding. 

2.  What  is  the  chief  use  of  injectors  in  stationary  power-plants?     Under  what  con- 
dition has  it  an  economic  advantage  over  other  kinds  of  feeders? 

3.  What  is  a  mechanically-driven  boiler-feed  pump?     A  motor-driven  boiler-feed  pump? 

4.  Why  is  mechanical  drive  ordinarily  more  efficient  than  electric  drive  for  a  boiler- 
feed  pump?      Demonstrate  with  an  example. 

5.  What  is  a  steam-driven  boiler-feed  pump?     What  operating  feature  of  a  pump  of 
this  type  gives  it  a  distinct  advantage  over  power  pumps? 


206  STEAM  POWER  PLANT  AUXILIARIES  [Div.  6 

6.  What  is  the  function  of  a  governor  on  a  direct-acting  steam  pump  in  boiler-feed 
service? 

7.  Describe  the  operation  of  a  diaphragm  type  of  pump  governor. 

8.  What  factors  mainly  decide  the  type  of  boiler-feed  pump  that  will  best  subserve 
the  economy  of  a  power  plant? 

9.  Why  are  centrifugal  pumps  generally  preferable  to  reciprocating  pumps  for  feeding 
boilers  of  installations  of  over  500  horsepower? 

10.  What  is  the  average  steam  consumption  of  steam-turbine-operated  boiler-feed 
pumps  in  plants  of  medium  capacity?     What  is  the  average  mechanical  efficiency  of 
these  pumps? 

11.  Why  are  power-pumps  better  adapted  than  steam  pumps  for  boiler-feeding  in 
non-condensing  power  plants  which  are  unequipped  with  heating  systems? 

12.  Why  will  downward  fluctuations  of  the  load  on  a  boiler  plant  impair  the  economy 
of  a  mechanically-driven  feed-pump  in  a  greater  ratio  than  in  the  case  of  a  steam- 
driven  feed-pump?     Demonstrate  with  an  example. 

13.  Why  should  both  steam-pumps  and  power-pumps  be  included  in  the  regular 
boiler-feed   equipment  of  a  non-condensing  plant  which  is  provided  with  an  extensive 
heating  system? 

14.  Describe  an  automatic  pumping  system  for  feeding  a  boiler  with  the  returns  from 
a  heating  system. 

15.  Describe  a  return-trap  system  of  boiler  feeding. 

16.  Explain  the  operation  of  a  Farnsworth  Duplex  Boiler  Feeder. 

17.  What  are  the  advantages  and  disadvantages  for  return  traps  as  compared  to 
pumps  for  boiler  feeding? 

18.  About  what  per  cent,  of  the  total  coal  is  used  indirectly  by  a  boiler  feed  pump 
in  a  well-designed  and  operated  plant? 

19.  About  what  per  cent,  of  the  exhaust  of  a  non-condensing  engine  is  necessary  to 
heat  the  feed  water? 

20.  What  is  meant  by  maintaining  an  exhaust-steam  "heat-balance"  in  a  power 
plant?     Describe  equipment  for  maintaining  such  a  heat-balance  automatically. 

21.  What  is  the  disadvantage  of  constant-speed  motors  for  feed-pump  drives?     What 
kind  of  motor  is  free  from  this  disadvantage? 

22.  What  are  two  disadvantages  of  centrifugal  pumps  as  stand-by  boiler-feeding 
equipment? 

23.  Why  should  reciprocating  power  pumps  be  fitted  with  relief  valves  under  some 
conditions? 

24.  What  are   the   two  general   methods   of   estimating   feed   water  requirements? 
Explain  each. 

25.  What  is  meant  by  "forcing"  a  boiler?     How  much  may  one  be  forced?     With 
what  results?     Under  what  conditions? 

26.  Name  four  common  troubles  of  pump  governors  and  give  their  remedies. 

PROBLEMS  ON  DIVISION  6 

1.  A  set  of  boilers  has  a  total  rating  of  600  boiler  h.p.     If  it  is  desired  to  have  a 
pump  capacity  of  50  Ib.  of  water  per  hr.  per  boiler  h.p.,  what  should  be  the  rating  of 
the  pump  in  gallons  per  hour.     If  it  is  later  decided  to  force  the  boilers  225  per  cent, 
at  peak  load,  what  capacity  should  the  pump  then  have  if  it  is  to  have  the  same  per 
cent,  excess  capacity  as  before? 

2.  The  main  engine  of  a  power  plant  has  a  duty  of  150  million  ft.  Ib.  per  1,000  Ib. 
of  steam  and  develops  500  h.p.     If  the  auxiliaries  require  10  per  cent,  as  much  steam 
as  the  main  engine  and  it  is  desired  to  have  a  feed  pump  capacity  50  per  cent,  in  excess 
of  normal  requirements,  how  many  gallons  per  hour  must  the  pump  deliver? 


DIVISION  7 
FEED -WATER  HEATERS 

239.  The  Reasons  That  Feed-Water  Heaters  Should  Be 
Used,  Fig.  218A,  are  these: 

(1)  If  cold  water  is  fed  into  a  boiler,  additional  fuel  must  be 
burned  to  raise  its   temperature  almost  to  the  boiling  point. 
This  represents  a  costly  waste  of  fuel,  inasmuch  as  in  practically 
every  plant  either  exhaust  steam  or  hot  flue  gases  or  both,  which 
would  otherwise  be  dissipated  into  the  atmosphere  and  lost,  can 
be  used  for  feed-water  heating. 

(2)  The  steel  plates  of  a  boiler  which  is  in  operation  are  very 
hot.     If  cold  water  is  fed  into  it,  certain  parts  of  the  boiler  shell 
may  thereby  be  cooled  excessively.     Thus  high  stresses  will  be 
produced  due  to  unequal  expansion  of  the  shell.     The  plates  are 
strained  as  are  also  the  riveted  joints.    Leakage  at  the  joints  and 
decreased  life  of  the  boiler  may  result. 

(3)  When  cold  water  is  pumped  into  boilers,  it  may  contain 
impurities,  which  tend  to  form  scale  on  the  inside  of  the  boilers 
when  it  becomes  hot.     This  scale  not  alone  interferes  with  the 
rate  of  transmission  of  heat  from  the  fire  to  the  water,  but  it  also 
may  permit  certain  parts  of  the  shell  to  become  excessively  hot 
because  the  water  in  the  boiler  is  prevented  by  the  scale,  from 
contacting    intimately    with    the    shell.     Blistering    and   short 
boiler  life  may  result.     But  if  the  water  is  first  heated  to  at  least 
200  deg.  fahr.  before  being  forced  into  the  boiler,  many  of  these 
impurities  may  be  thereby  precipitated  in  an  external  chamber, 
from  which  they  can  be  removed  readily.     Thus  they  are  prevented 
from  entering  the  boiler. 

240.  In  A  Non-Condensing  Plant  Eighty  Per  Cent.  Of  The 
Energy  In  The  Live  Steam  Is  Wasted  In  The  Exhaust.— That 
is,  the  amount  of  heat  remaining  in  the  exhaust  steam  from  a 
non-condensing  engine  is  about  80  per  cent,  of  the  original 
heat  imparted  in  the  boiler  to  the  steam.     The  truth  of  this 
may  be  shown  thus: 

207 


208 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


SBC.  240]  FEED-WATER  HEATERS  209 

EXAMPLE. — Consider  a  medium-capacity,  well-maintained  non-con- 
densing plant  operating  at  150  Ib.  per  sq.  in.  boiler  pressure.  Such  a 
plant  should  develop  an  indicated  horsepower  hour  (i.h.p.  hr.)  on  25  Ib. 
of  steam.  That  is,  its  water  rate  would  be  25  Ib.  of  steam  per  i.h.p.  hr. 
Assume  that  the  cold  feed  water  has  a  temperature  of  50  deg.  fahr. 
Hence,  we  are  interested  only  in  the  heat  which  must  be  added  to  this 
cold  feed  water  to  raise  it  to  the  temperature  of  steam  at  150  Ib.  per  sq.  in. 

From  a  steam  table  it  is  found  that  the  heat  which  must  be  added  to 
1  Ib.  of  water  at  50  deg.  fahr.  to  convert  it  into  steam  at  150  Ib.  pressure 
is  1,177  B.t.u.  Hence,  on  this  basis  the  25  Ib.  of  steam  which  is  required 
by  the  engine  to  produce  1  h.p.  hr.  represents:  25  X  1,177  =  29,425 
B.t.u.  Now,  from  a  conversion  table,  it  is  found  that  1  h.p.  hr.  is  equal 
to  2,545  B.t.u.  Therefore,  out  of  the  29,425  B.t.u.  imparted  to  each 
pound  of  steam,  only  2,545  B.t.u.  is  converted  into  useful  work  in  the 
production  of  1  h.p.hr.  Thus  there  must  be  in  the  exhaust  steam  from 
the  engine  (disregarding  radiation):  29,425  -  2,545  -=  26,880  B.t.u.  per 
i.h.p.  hr.  The  percentage  of  heat  converted  into  work  on  the  engine 
piston  must,  then,  be:  2,545  -J-  29,425  =  0.087  =  8.7  per  cent. 

If  the  radiation  losses  are  assumed  to  be  10  per  cent,  (of  the  heat  in 
the  exhaust  steam)  which  is  a  fair  average  value,  the  available  heat  per 
indicated  h.p.  hr.  would  be:  26,880  --  2,688  =  24,192  B.t.u.  The 
percentage  of  the  total  heat  received  by  the  engine  which  is  lost  in  radi- 
ation is:  2,688  -5-  29,425  =  0.091  =  9.1  per  cent.  Hence  the  percentage 
(of  the  original  heat  which  was  in  each  pound  of  steam)  that  is  now 
available  in  the  exhaust  is:  24,192  -r-  29,425  =  0.822  =  82.2  per  cent. 
Note,  then,  that  about  82  per  cent,  of  the  original  heat  is  available  in 
the  exhaust  steam  from  the  engine.  Thus,  summarizing,  the  percen- 
tages of  the  heat  units  delivered  to  the  engine  cylinder  in  the  live  steam 
are  either  expended  or  available  thus: 

Heat  expended  as  work  on  engine  piston 8.7  per  cent. 

Heat  lost  in  radiation 9 . 1  per  cent. 

Heat  available  in  exhaust  steam 82 . 2  per  cent. 

Total 100 . 0  per  cent. 

Where  larger  engines  and  turbines  operating  condensing  and  with 
superheat  are  used,  a  greater  proportion  of  the  heat  is  realized  in  useful 
work.  A  non-condensing  prime  mover  discharges  its  exhaust  steam  into 
the  atmosphere  at  212  deg.  fahr.  A  condensing  prime  mover  discharges 
its  exhaust  steam  into  its  condenser  at  a  temperature  of  about  100  deg. 
fahr.  or  lower,  depending  on  the  vacuum  maintained.  Even  with  the 
most-efficient,  condensing,  steam-power-plant  equipment,  where  the 
water  rate  is  as  low  as  10  Ib.  of  steam  per  h.p.  hr.,  about  75  per  cent,  of 
the  heat  is  discharged  with  the  engine  exhaust  and  is,  for  all  practical 
purposes,  lost. 

14 


210 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


241.  The  Two  General  Types  Of  Feed-Water  Heating 
Equipment  are :  (A)  Exhaust  steam  feed-water  heaters  (which 
are  treated  in  this  Division)  which  are  devices  which  use  the 
exhaust  steam  for  raising  the  temperature  of  the  feed-water. 

(B)  Economizers  (Div.  8)  which 
are  devices  which  use,  for  heat- 
ing the  feed  water,  the  hot  flue 
gases  after  they  are  discharged 
from  the  boiler-furnace. 

242.  Exhaust  Steam  Feed- 
Water  Heaters  are  of  many 
types  but  may  be  classified  into 
two  general  divisions:  (1)  The 
open  heater,  Fig.  219.  (2)  The 
closed  heater,  Figs.  220  and  221. 


^-7b  Feed  Pump    '-Hot  Water 


FIG.  219.— Diagram  of  Open  Feed-  gy  an  open  heater  is  meant  one 

Water  Heater.  .    J  . ^ 

in  which   the  exhaust  steam  is 

permitted  to  contact  directly  in  a  suitable  chamber  with 
the  cold  water  which  is  to  be  heated.  Thus  part  of  the 
exhaust  steam  is  condensed  in  raising  the  temperature  of 
the  cold  water  and  is  used  as  part  of  the  feed  water.  With 


I    Safely  Wye 
Feed  Outlet   Connection-^ 


Brass  Water  Tubes-' 
—Floating  Tube  Head 

C<J  B<J 

FIG.  220. — "  Blake-Knowles"  Water-Tube  Type  Of  Closed  Exhaust-Steam  Feed- 
Water  Heater.  (The  water  passes  through  each  of  the  six  tube  nests  in  turn, 
thus  traversing  the  heater  six  times.  The  steam  passes  three  times  through  the 
heater.) 

an  open  heater,  the  temperature  of  the  feed  water  can — 
assuming  that  sufficient  exhaust  steam  is  available,  and  there 
usually  is,  be  raised  to  a  temperature  of  210  to  212  deg.  By 


SEC.  243] 


FEED-WATER  HEATERS 


211 


a  closed  heater  is  meant  one  in  which  the  steam  does  not  contact 
with  the  cold  water  but  in  which  the  heat  from  the  exhaust 
is  imparted  to  the  water  through  the  walls  of  tubes.  Thus 
in  the  closed  type,  the  water  to  be  heated  and  the  exhaust 
steam  for  heating  it  are  confined  to  separate  chambers. 

NOTE. — THE  CLOSED  HEATER  MUST 
BE  USED  WHERE  THE  BOILER  FEED- 
WATER  MUST  BE  MAINTAINED  ABSO- 
LUTELY FREE  FROM  OIL.  An  oil 
separator,  which  extracts  practically  all 
of  the  oil,  always  forms  a  part  of  open 
heater  equipments,  but  these  separators 
cannot  always  be  relied  upon  to  extract 
all  of  the  oil  from  the  exhaust  steam. 


'Shell 


.-Heotoler 


Wafer  From  Feed  Pump*.. 


m 


-OirfetlbBoikr 


Drain  For- ••• 
Conofensate 


FIG.  221. — Diagram   Of   Closed 
Feed- Water  Heater. 


243.  Economies  Accruing  Due 
To  The  Use  Of  Feed-Water 
Heaters  are  very  pronounced.  In 
the  average  plant  a  saving  of  from 
11  to  14  per  cent,  in  fuel  may  be 
expected  due  to  the  installation  of 
a  heater.  There  is  usually  sufficient 
exhaust  steam  (see  Sec.  209)  which  would  otherwise  be  wasted, 
available  to  heat  the  feed  water.  All  of  the  heat  which  can 
be  imparted  to  the  feed  water  before  it  is  pumped  into  the 
boiler  represents  that  much  saving  in  fuel.  A  temperature  of 
212  deg.  fahr.  is  the  highest  to  which  water  can  be  raised  (at 
atmospheric  pressure)  without  its  being  converted  into  steam. 
It  follows  that  every  effort  should  be  made  to  utilize  exhaust 
steam  to  raise  the  feed  water  to  212  deg.  fahr.  While  a  tem- 
perature of  212  deg.  fahr.  may  not  be  feasible  in  every  case,  it 
is  usually  possible  to  attain  a  feed- water  temperature  of  210 
or  211  deg.  fahr.  Because  a  higher  feed-water  temperature 
can  be  obtained  with  an  open  heater  than  with  a  closed  one, 
the  open  type  is  somewhat  more  economical.  Every  steam- 
power  plant  should  have  a  feed-water  heater. 

NOTE. — THE  FOLLOWING  RULES  FOR  ESTIMATING  THE  APPROXIMATE 
FUEL  SAVING  DUE  To  PREHEATING  FEED- WATER  are  often  useful: 
(1)  For  every  11  deg.  fahr.  which  is  added  to  the  temperature  of  the  feed- 
water  with  exhaust  steam  there  results  a  saving  of  about  1  per  cent,  of  the 
fuel  which  would  otherwise  be  required.  (2)  For  a  given  consumption  of 


212  STEAM  POWER  PLANT  AUXILIARIES  [Drv.  7 

fuel,  the  evaporative  capacity  of  a  boiler  is  increased  by  approximately  1 
per  cent,  for  each  11  deg.  fahr.  increase  of  the  feed-water  temperature. 

EXPLANATION.  —  Suppose  the  temperature  of  the  feed-water  is  60  deg. 
fahr.,  and  the  boiler  pressure  is  120  Ib.  per  sq.  in.,  gage.  According  to 
the  steam  tables,  found  in  any  engineering  handbook,  the  total  heat, 
above  32  deg.  fahr.,  of  steam  at  120  Ib.  per  sq.  in.,  gage,  is  1191.6  B.t.u. 
per  pound.  Therefore,  the  total  heat  that  must  be  supplied  to  each  pound 
of  the  feed-water  is  [1191.6  -  (60  -  32)]  =  1163.6  B.t.u.  If,  now,  the 
feed-water  temperature  is  raised  to  71  deg.  fahr.  by  waste  heat,  the 
saving  =  (71  —  60)  =11  B.t.u.  per  pound.  Then  the  per  cent,  saving  = 
11  -T-  1163.6  =  0.009,5  or  roughly  1  per  cent,  of  the  total  heat  supplied  to  the 
steam. 

EXAMPLE.  —  A  power  plant,  in  which  the  boilers  develop  1,000  boiler 
h.p.  with  feed  water  at  100  deg.  fahr.,  is  furnished  with  a  heater  which 
supplies  the  feed  water  at  210  deg.  fahr.  What  additional  boiler  horse- 
power is  thus  realized? 

SOLUTION.  —  By  Sec.  243  the  evaporative  capacity  of  the  boilers  is 
increased  approximately  1  per  cent,  for  each  11  deg.  fahr.  increase  of  the 
feed-water  temperature.  Hence,  the  power  of  the  boilers  is  increased 


244.  The  Saving  Of  Heat  Which  Results  From  Preheating 
Boiler  Feed-Water  with  exhaust  steam  that  would  otherwise 
be  wasted  may  be  computed  by  the  following  formula: 

(76)  H,=  ~  100  (per   cent.) 


Wherein  H/  =  the  saving,  in  per  cent,  of  the  heat-content  of 
the  fuel.  T/i  =  the  temperature  of  the  feed-water,  in  degrees 
Fahrenheit,  before  preheating.  T/2  =  the  temperature  of 
the  feed-water,  in  degrees  Fahrenheit,  after  preheating. 
H  =  the  total  heat  in  the  steam  which  is  generated  in  the 
boiler,  in  British  thermal  units  per  pound. 

NOTE.  —  The  specific  heat  of  water  varies  somewhat  with  the  tempera- 
ture (see  the  author's  PRACTICAL  HEAT).  In  the  compilation  of  For. 
(76),  however,  the  specific  heat  of  the  feed-water  is  assumed  to  have  a 
constant  value  of  1.0  B.t.u.  per  Ib.  for  all  temperatures.  Computations 
based  upon  this  assumption  are  correct  within  1  per  cent.,  which  is  suf- 
ficiently accurate  for  all  practical  purposes. 

EXAMPLE.  —  A  boiler  generates  steam  at  a  pressure  of  100  Ib.  per  sq. 
in.,  gage.  The  water  which  is  fed  to  the  boiler  is  preheated,  with  exhaust 
steam,  from  80  deg.  fahr.  to  210  deg.  fahr.  What  saving  of  heat  results 
from  thus  utilizing  the  exhaust  steam? 


SEC.  245] 


FEED-WATER  HEATERS 


213 


SOLUTION. — As  given  in  a  table  of  the  properties  of  saturated  steam, 
the  total  heat  in  steam  at  100  Ib.  per  sq.  in.,  gage,  is  1188  B.t.u.  per  Ib. 
Hence,  by  For.  (76),  the  saving  =  Hf  =  { (T/2  -  Tfl)/[H  -  (Tfl  - 
32)]}  100  =  {(210  -  80)  +  [1188  -  (80  -  32)]}  X  100  =  11.4  per  cent. 

245.  The  Percentage  Of  Fuel  Saving  Due  To  Feed-Water 
Heating  May  Be  Computed  Graphically  (Fig.  222)  for  satu- 
rated or  superheated  steam.  Points  A  and  C,  for  instance, 
are  found  corresponding  to  initial  and  final  feed-water  tem- 
peratures. A  vertical  line  from  A  is  traced  until  it  intersects 
an  oblique  line  from  C  at  B.  A  point  D  is  then  found  on  the 


m  rressurc 
5q.  In  Gage 

'./ 

Sorfura/feo/  Steam--''' 

Temp.Vtotcr  leaving  HectteF 


Initial  Feed-Water.  Terrperotfure  Deg.Fahr: 
FIG.  222.— Graph  Showing  Percentage  Of  Fuel  Saved  By  Heating  Feed  Water. 

scale  at  the  upper  left  corresponding  to  the  steam  gage  pres- 
sure. A  vertical  line  from  D  is  traced  to  its  intersection  with 
a  graph  for  saturated  steam  at  E  or  some  degree  of  superheat 
at  F.  Lines  from  F  and  E  are  traced  horizontally  to  the  line 
A  B  and  then  obliquely  until  they  intersect  a  horizontal  line 
from  B  at  G  and  H.  The  saving  in  each  case  may  be  read 
from  the  per  cent,  scale. 

EXAMPLE. — In  the  case  selected  in  Fig.  222  the  initial  temperature 
was,  A,  110  deg.  fahr.  The  water  left  the  heater  at  C,  210  deg.  fahr. 
The  gage  pressure  was,  D,  160  Ib.  per  sq.  in.  For  saturated  steam,  the 
saving  was,  G,  9.1  per  cent.  For  100  deg.  fahr.  superheat  the  saving 
was,  H,  8.7  per  cent. 

246.  The  Net  Monetary  Saving  Which  Results  From 
Preheating  Boiler  Feed-Water  With  Exhaust  Steam  that 
would  otherwise  be  wasted  must  be  computed  upon  a  basis 


214  STEAM  POWER  PLANT  AUXILIARIES  [Drv.  7 

of  the  interest  on  the  investment  in  heating  apparatus  and  the 
annual  cost  of  depreciation,  attendance  and  maintenance, 
taken  in  conjunction  with  the  annual  heat-saving  effected, 
which  may  be  computed  by  using  For.  (76). 


EXAMPLE. — The  coal-consumption  of  a  battery  of  boilers  which  receive 
feed-water  at  a  temperature  of  110  deg.  fahr.  is  3  tons  per  day.  It  is 
estimated  that  by  utilizing  a  quantity  of  exhaust  steam  which  is  now 
going  to  waste,  the  feed-water  may  be  preheated  to  212  deg.  fahr.  The 
average  steam-pressure  is  110  Ib.  per  sq.  in.,  gage.  The  coal  costs  $3 
per  ton.  The  plant  operates  310  days  per  year.  The  cost  of  the  feed- 
water-heating  apparatus  and  its  installation  will  be  $300.  The  rate  of 
interest  on  the  investment  is  6  per  cent,  per  annum.  The  assumed  rate 
of  depreciation  is  5.0  per  cent,  per  annum.  The  cost  of  maintaining  and 
operating  the  apparatus  is  presumed  to  be  $5  per  month.  What  will  be 
the  probable  annual  net  saving? 

SOLUTION. — As  given  in  a  table  of  the  properties  of  saturated  steam, 
the  total  heat  in  steam  at  a  pressure  of  110  Ib.  per  sq.  in.,  gage,  is  approxi- 
mately 1,190  B.t.u.  per  Ib.  Hence,  by  For.  (76),  the  probable  thermal 
saving  =  Hf  =  { (Tfz  -  Tfl)/[H  -  (Tfl  -  32)]}  100  =  {[(212  -  110)  -r- 
[1,190  -  (110  -  32)]}  X  100  =  9.17  per  cent.  The  present  annual  cost 
of  the  coal  supply  =  3X3X3 10=  $2,790.  Therefore,  the  probable  re- 
duction in  the  annual  coal  bill,  due  to  utilizing  the  available  exhaust 
steam  =  (2,790  X  9.17)  -r-  100  =  $255.84.  The  interest  on  the  invest- 
ment =  (300  X6)  -r-  100  =$18.  The  annual  cost  of  depreciation  = 
(300  X  5.0)  -s-  100  =  $15.00.  The  annual  cost  of  maintenance  and  oper- 
ation =  (12  X  5)  =  $60.  Hence,  the  approximate  net  annual  saving  will 
be  255.84  -  (18  +  15  +  60)  =  $162.84. 

In  other  words  the  feed-water-heating  equipment  will  pay  for  itself  in 
about  2  yr.  If  the  heater  were  installed  in  a  plant  where  it  would  not 
be  necessary  to  employ  additional  labor  to  maintain  it,  it  would  pay  for 
itself  in  about  \Y%  yr. 

NOTE. — OP  ALL  BOILER  ROOM  ACCESSORIES,  FEED- WATER  HEATERS 
ARE,  PROBABLY,  THE  MOST  EFFECTIVE  SAVERS  OF  COAL.  (From  the 
AMERICAN  CORRESPONDENCE  SCHOOL.)  With  condensing  engines,  the 
condensate-pump  discharges  from  the  condenser  into  the  hot  well.  Then 
the  water  is  drawn  from  the  hot  well  as  boiler  feed  at  a  temperature  of 
100  deg.  to  140  deg.  F.  This,  however,  if  the  boiler  pressure  is  over 
100  Ib.  per  sq.  in.,  is  not  a  sufficiently-high  temperature  for  the  best 
economy.  Feed  water  at  this  temperature  should  be  passed  through  a 
feed-water  heater.  With  non-condensing  engines  it  is,  from  a  standpoint 
of  economics  absolutely  necessary  that  in  some  way  the  feed  water  be 
heated  by  the  exhaust  steam  in  a  feed-water  heater  or  by  the  waste  gases 
from  the  chimney  in  an  economizer. 


SEC.  247] 


FEED-WATER  HEATERS 


215 


247.  Exhaust-Steam  Feed-Water  Heaters  May  Be  Classi- 
fied With  Respect  To  Their  Relation  To  Other  Plant  Equip- 
ment (see  also  Sec.  242) ,  as  hereinafter  explained,  as  primary 
and  secondary  heaters.     They  may  be  classified  with  respect  to 
the  steam  pressure  used  as  atmospheric,  vacuum  and  pressure 
heaters. 

248.  Table     Showing     Classification    Of    Representative 
American   Feed-Water   Heaters    (partly    from    Gebhardt). 


Exhaust 
steam 

Open 
Atmospheric 

Bonar 
Blake-Knowles 
Cochrane 
Cookson 
Elliot 
Hoppes 

Moffat 
Reliance 
Sims 
Stillwell 
Webster 

Vacuum,  pressure, 
or     atmospheric 
^        (water  tube). 

o> 

02 

3 

American 
Griscom  Russel 
Gaubert 
National 

Ross 
Standard 
Wainwright 
Wheeler 

u  • 

Vacuum,  pressure, 
or     atmospheric 
(steam  tube). 

Berryman 
Kelly 

Otis 
Ross 

Live  steam 

Open  pressure 

Hoppes 
Baragwanath 

249.  A  Primary  Or  Vacuum  Heater  is  a  closed  feed-water 
heater  which  is  connected  to  the  exhaust  of  a  condensing  engine 
between  the  engine  and  the  condenser.  The  conditions  favor- 
able to  the  installation  of  a  primary  heater  exist  where  the 
supply  of  exhaust  steam  from  the  auxiliaries  in  a  condensing 
plant  is  insufficient  for  properly  heating  the  feed  water.  In 
such  cases  (Fig.  223)  the  feed-water  can  first  be  heated  in  the 
primary  heater,  with  steam  exhausting  from  the  engine  and 
then  be  passed  through  a  secondary  heater  (Sec.  250)  which  is 
supplied  with  exhaust  steam,  at  atmospheric  pressure  or  above, 
from  the  auxiliaries.  The  primary  heater  is  under  about  the 
same  vacuum  as  the  condenser.  If  the  condenser  maintains 
a  vacuum  of  26  in.,  the  temperature  of  the  discharge  from  the 


216 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


primary  heater  will  probably  not  exceed  118  deg.  fahr.  The 
primary  heater  also  acts  as  a  supplementary  surface  condenser 
in  which  the  feed-water  acts  as  condensing  water. 

NOTE. — A  PRIMARY  HEATER  Is  ESPECIALLY  USEFUL  WHERE  A  JET 
CONDENSER  Is  USED  AND  THE  CONDENSER  WATER  Is  UNSUITED  FOR 
BOILER  FEED.  When  this  is  true,  fresh  water  must  be  used  as  boiler 
feed  and  is  usually  supplied  at  much  below  hot-well  temperatures.  For 
instance,  if  an  average  hot-well  temperature  is  100  deg.  fahr.  and  the 


--Siphon 
Conol&nser 


Vacuum  Pipe 
,,-To  Trap 
•'  Atmospheric, 
Or  Szcono/o/ry 
Hzoitzr— 


?:  v*.r-£?u'r; 


FIG.  223.— Showing  Method  Of  Installing  Primary  And  Secondary  Feed-Water  Heaters. 

water  supply  temperature  is  60  deg.  fahr.,  additional  heater  capacity  is 
necessary  to  raise  the  water  the  difference  of  40  deg.  fahr.  The  same 
necessity  for  a  primary  heater  exists  when  a  high  vacuum  is  obtained 
with  a  surface  condenser.  The  condensate  may  then  be  cooled  to  60 
deg.  fahr.  or  a  lower  temperature. 

250.  An  Atmospheric  Heater  is  an  open  or  closed  feed- 
water  heater  (Fig.  224)  which  utilizes  the  exhaust  from  non- 
condensing  engines  or  auxiliaries.  The  pressure  on  these  heaters 
is  equal  to  the  back-pressure  on  the  engines  which  supply 
the  exhaust  steam.  Where  auxiliaries  supply  the  exhaust, 
this  pressure  is  usually  controlled  by  a  back  pressure  valve,  at 


SKC.  251] 


FEED-WATER  HEATERS 


217 


a  few  pounds  above  atmospheric.  Where  the  exhaust  from 
the  heater  is  used  in  a  vacuum  heating  system,  the  pressure 
may  be  a  few  inches  mercury  column  below  atmospheric.  The 
maximum  feed-water  temperatures  obtainable  in  atmospheric 
heaters  are  about  200  deg.  fahr.  in  closed  heaters  and  210  deg. 
fahr.  in  open  heaters. 


5 f earn  Supply  Pipe 
Non-Condensing  Engine^ 


Delivery  Pipe  To  Boilers-,         ^ 


'-Boiler  Feed  Pump       '-Through  Type  Of  Exhctusf-Sieotm  Feed-Water  Heater 

Fid.  224. — Showing  Piping  Arrangement  Of  Stilwell  Through  Type  Of  Exhaust-Steam 
Feed-Water  Heater  In  Non-Condensing  Plant. 

251.  Both  Vacuum  And  Atmospheric  Heaters  May  Be 
Used  In  Condensing  Plants  (Fig.  223). — The  feed-water  is 
first  forced  by  the  feed  pumps  through  the  vacuum  heater,  in 
which  it  absorbs  whatever  heat  may  be  abstracted  from  the 
exhaust  steam  coming  from  the  main  engine.  The  feed-water 
then  passes  through  the  atmospheric  heater  and  on  to  the 
boilers.  Exhaust  steam  from  the  pumps  or  from  any  other 
source,  which  it  may  be  inconvenient  or  unprofitable  to  con- 
dense, is  piped  to  the  atmospheric  heater.  When  an  atmos- 
pheric heater  is  connected  in  this  way  it  is  commonly  called  a 
secondary  heater  as  distinguished  from  a  primary  or  vacuum 
heater. 

NOTE. — THE  SECONDARY  HEATER  MAY  BE  OF  THE  OPEN  OR  CLOSED 
TYPE.  When,  however,  it  is  of  the  open  type,  the  feed  water  must  flow 
by  gravity — or  be  forced  by  a  separate  pump — through  the  primary 


218 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


heater.     It  is  usual  therefore  to  select  secondary  heaters  of  the  closed 
type.     The  primary  heater  is  always  of  the  closed  type. 

252.  Installation  Of  Primary  And  Secondary  Feed-Water 
Heaters,  To  Be  Operated  Alternately  (Fig.  225),  may  be  advis- 
able for  condensing  plants  in  which  the  quantity  of  exhaust 
steam  from  the  auxiliaries  is,  ordinarily,  sufficient  for  feed- 
water  heating,  but  where  the  condenser  auxiliaries  are  occa- 


feed-&/mp  Sucfhri-'         '^team-Supply 


Fia.  225.—  Installation  Of  Primary  And  Secondary  Heaters  For  Alternate  Operation. 

sionally  inoperative  on  account  of  the  main  engine  being 
required  to  exhaust  to  the  atmosphere.  With  such  installa- 
tions the  primary  heater  can  be  used  alone  at  such  times  as  the 
main  engine  is  running  non-condensing,  while  the  secondary 
heater  can  be  used  alone  when  the  operation  is  condensing. 

253.  The  Back-Pressure  On  An  Engine  May  Not  Be 
Increased  By  Installing  A  Feed  -Water  Heater  in  the  exhaust 
line.  This  is,  with  closed  heaters,  due  to  the  fact  that  the 
shell  of  the  heater,  in  the  case  of  a  water-tube  heater,  or  the 


SEC.  253] 


FEED-WATER  HEATERS 


219 


nest  of  tubes,  in  the  case  of  a  steam-tube  heater,  is,  usually, 
of  much  greater  cross-sectional  area  than  the  exhaust  pipe. 
Also,  the  partial  condensation  of  the  exhaust  steam,  due  to 
absorption  of  heat  therefrom  by  the  feed-water,  tends  to 


To  Atmosphere 

Mult /port  Valve >. 

Thersnosfaf- 


Exhaust  To  Heating  A  no/  Dry  mo/  Coifs-, 
'•Co/of  Wafer  Supply 


Conafensafe  Her  urns 
Hot  feed-Wafer  To  Bo/'/ers^ 
Bo/'/er  Feed 


'     uyfrfihy  Circuits  \ 
ligrhf/ng  Company's  Street  Circuit-' 

b/ne-Dr/Yen  Inofuction 
Generator 


FIG.  226. — Cochrane  Open  Induction  Heater,  H,  Equipped  With  Automatic  Thermo- 

static  Valve,  Used  For  Exhaust  Steam-Heating  System. 

(In  the  ordinary  power  plant  which  uses  exhaust  steam-heating,  the  power  and  heating 
requirements  rarely  balance.  Some  of  the  time,  perhaps  half  the  year,  steam  is  wasted 
to  atmosphere.  On  the  other  hand,  central-station  energy  may  be  used  when  more 
power  is  required  than  the  heating  system  will  generate  as  a  by-product.  Or,  an  auto- 
matic heat  balance  for  such  conditions  may  be  provided  by  the  arrangement  shown 
above.  The  back-pressure  turbine,  T,  exhausts  into  a  steam-stack  heater,  H,  and  to 
the  heating  system,  <S.  The  generator,  G,  supplies  energy  to  the  local  circuit,  which  is 
also  connected  to  the  central-station  company's  street  mains  through  a  meter. 

A  thermostat,  responsive  to  the  temperature  of  the  water  in  the  heater,  governs  the 
admission  of  steam  to  the  turbine  (subject,  of  course,  to  an  automatic  speed  limit). 
When  the  power  requirements  are  greater  than  the  heat  requirements,  central-station 
energy  is  taken  through  the  meter.  If  at  times  more  heat  than  power  is  required,  steam 
can  be  by-passed  automatically  or  power  can  be  sold  back  to  the  electric  company. 
The  conversion  of  heat  to  mechanical  power  and  building  heating  is,  with  this  arrange- 
ment, practically  100  per  cent,  perfect.  No  heat  is  wasted  to  atmosphere  or  to  con- 
denser circulating  water.) 

prevent  back-pressure.  An  open  induction  heater  (Sec.  254) 
with  an  extra-large  oil-separator  may  be  used  on  a  non-con- 
densing engine  exhaust  without  increasing  the  back-pressure 


220 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


more  than  %  Ib.  per  sq.  in.  Where  the  engine  exhaust  is 
used  for  feed- water  heating  only,  an  open  heater  arranged 
thus  and  properly  vented  and  managed  will  heat  the  feed 
water  to  within  about  2  deg.  fahr.  of  the  exhaust  steam 
temperature.  Some  open  heater  manufacturers  claim  that 
an  open  heater  need  not  cause  any  additional  back-pressure. 

NOTE. — A  BACK-PRESSURE  VALVE  INCREASES  THE  EFFECTIVENESS  OF 
A  FEED-WATER  HEATER  and  acts  also  as  a  safety  valve  for  the  heater. 
It  should  be  a  reliable  easy-moving  valve  of  large  port  opening.  When 
an  induction  heater  and  a  heating  system  are  supplied  with  steam  from 
the  same  exhaust  line  (Fig.  226),  a  back-pressure  valve  is  necessary  to 
insure  proper  distribution  of  the  steam  to  all  the  heating  equipment. 
A  back-pressure  valve  decreases  the  power  developed  by  an  engine  about 
2}^  per  cent,  for  each  pound  of  back  pressure.  The  cost  of  the  decreased 
engine  efficiency  due  to  back  pressure  carried  for  a  heating  system  is 
usually  much  less  than  the  cost  of  the  live  steam  which  would  be  required 
for  heating  if  the  back  pressure  were  not  maintained.  The  decrease  in 
engine  power  due  to  a  back  pressure  may  be  made  up  by  carrying  2  to  5 
Ib.  per  sq.  in.  greater  boiler  pressure  for  each  pound  of  back  pressure — 
which  will  of  course  require  the  burning  of  a  slightly  greater  amount 
of  coal. 

254.  Exhaust-Steam  Feed-Water  Heaters  May  Be  Classi- 
fied According  To  Their  Piping  Arrangements  as:  (1)  Induced 

or  draw  heaters  (Figs.  227,  228, 
229,  230,  231,  232),  which  re- 
ceive no  more  exhaust  steam 
from  the  available  supply  than 
the  water  will  entirely  condense. 
(2)  Through  or  thoroughfare  heat- 
ers (Figs.  233  and  234),  which 
receive  all  of  the  available 
supply  of  exhaust  steam.  With 
the  first  arrangement,  complete 
condensation  of  the  steam 
which  passes  into  the  heater 
induces  a  continual  flow  thereto 
through  a  branch  from  the  main 
exhaust  pipe.  If  the  quantity 
of  steam  exhausted  by  the  engine  is  greater  than  that  which 
can  be  condensed  in  the  heater,  the  excess,  with  the  first 


Exhaust  To  Atmosphere 


Exhaust 
From 


Engine^ 


.\>\\\\\\\\\\\\\\\ 
Drai'n--- 


\\Y\\V\\\\\ 

Water  Outlet-''        r     i 
Drain  Connection^' 


FIG.  227. — Horizontal  Closed  Heater 
Piped  For  Service  On  The  Induction 
Principle. 


SEC.  254] 


FEED-WATER  HEATERS 


221 


arrangement,  may  go  directly  from  the  engine  to  the  atmos- 
phere, or  to  a  heating  system  or  condenser.     With  the  second 

-ff/'sers  To  Heating  Si/sfem^ 


..--Back-Pressure 

Valve 

.Wcttzr-Supply  P/P& 
'-Steam  Main 


Horizontal  Heater— 
'zed-Pump  Suction'-., 


Fia.  228. — Hoppes    Horizontal    Exhaust-Steam    Feed-Water    Heater    Installed    For 
Induction  Operation  With  Gravity  Heating  System. 


Air  tonW 


CZI 


Steam 
Supply 


-Exhaust  Steam  To  Atmosphere 
-Backpressure  Valve        High  Pressure 
Steam  Main-— 


Hlgh  Pressure,  Drips 
•Feed  Water  rkater 
.--Air  Vain 

k  .Vacuum  Brea 
JJ    Check  Valve 


Fia.  229.— Typical   Installation   Of   Open   Feed-Water   Heater   Pipes   For   Induction 
Operation  In  Connection  With  Vacuum  Heating  Plant. 

arrangement,  if  more  steam  is  received  than  can  be  condensed 
in  the  heater,  the  excess  passes  through  the  heater  to  the 
atmosphere  or  heating  system. 


222 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


NOTE. — IF  THE  QUANTITY  OF  EXHAUST  STEAM  AVAILABLE  FOR  HEAT- 
ING THE  FEED-WATER  Is  EXCESSIVE,  the  open  heater  should  (Fig.  228) 
be  arranged  for  induction  service.  If  the  surplus  exhaust  steam  from 


•To  Heating  System  Or 
To  Atmosphere 


^Impulse  Of  Steam  Current 


FIG.  230.—  Impulse  Of  Steam  Current  Di- 
rected Toward  Induction  Heater. 


FIG.  231. — Impulse  Of  Steam  Current  At 
Right  Angles  To  Induction  Heater. 


•Back-Pressure  Valve  r1.-f?eceiver  Separators 


FIG.  232.— Induction-Type  Open  Feed- Water  Heater,  H,  Installed  In  Connection 
With  Reciprocating  Engine,  E,  And  Mixed-Flow  Turbine,  T.  (Exhaust  from  recipro- 
cating-engine, feed-pump  turbine,  F,  and  auxiliary  turbine,  A,  piped  to  feed-water 
heater,  heating  system  and  mixed-flow  turbine.) 

an  induction  heater  is  used  for  heating  or  drying  purposes,  and  the  result- 
ing condensate  is  afterwards  returned  to  the  heater,  the  surplus  steam 


SEC.  255] 


FEED-WATER  HEATERS 


223 


should  be  passed  through  an  independent  oil  separator.  With  induction 
operation  the  surplus  steam  will  pass  on  in  a  much  drier  condition  than 
if  it  had  gone  through  the  heater.  If  the  condensate  from  a  closed  heater 


Condenser 


FIG.  233. — Showing  Piping  Arrangement  Of  Stilwell  Through-Type  Of  Exhaust-Steam 
Feed- Water  Heater  In  Condensing  Plant. 


TO  Boilers, 


Closed 
Feed-Water 
Heater-—. 


•r-To  Hunting  System  Or  Atmosphere 


FIG.  234.— Equipment  Of  Closed  Feed-Water  Heater  Installed  For  Service    On  The 
Thoroughfare  Principle. 

is  to  be  returned  to  an  open  heater,  the  inlet  to  the  closed  heater  should 
be  fitted  with  an  oil  separator. 

255.  The  Piping  Of  An  Induction  Heater  should  be  so 
arranged,  when  possible,  that  the  direct  impulse  of  the  exhaust- 
steam  current  (Fig.  230)  is  toward  the  heater,  rather  than 


224 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


Exhaust  Outlet- 
V® 


$fw/  She//, 


toward  the  atmosphere  or  heating  system  (Fig.  231).  The 
object  of  this  is  to  insure  delivery  to  the  heater  of  as  much 
steam  as  it  can  accommodate.  With  the  impulse  of  the 
steam  current  at  right  angles  to  the  heater  (Fig.  231)  the 
heater  might  receive  a  scanty  or  starved  supply. 

256.  The  Temperature  Of  The  Exhaust  Steam  Entering  A 
Feed-Water  Heater  depends  upon  the  back-pressure.  If  the 
steam  in  excess  of  that  which  is  condensed  in  the  heater  is 

discharged  directly  to  the 
atmosphere,  then  the  back 
pressure  is,  ordinarily,  at- 
mospheric pressure. 
Hence,  in  such  cases  the 
temperature  is  about  212 
deg.  fahr.  But  if  the  ex- 
cess of  steam  is  used  in  a 
heating  system,  the  back 
pressure  may  range  from 
atmospheric  up  to  about  5 
Ib.  per  sq.  in.  In  the  lat- 
ter case  the  temperature 
would  be  about  227  deg. 
fahr. 

257.  Open  Exhaust- 
Steam  Feed-Water  Heat- 
ers Are  Generally  Designed 
To  Perform  A  Four-Fold 
Function  as  follows :  (I)  To 
remove  the  oil  from  the  ex- 
haust steam  which  supplies 
the  heat.  This  is  accom- 
FIG.  235.— The  Moffat  Open  Exhaust-Steam  plished  by  means  of  an 

Feed- Water  Heater  And  Purifier.  oil-Separating    device    (Fig. 

236)  which  (Fig.  235)  usually  forms  an  integral  part  of 
the  heating  apparatus.  (2)  To  bring  the  exhaust  steam  and 
feed-water  into  intimate  contact.  The  heating  effectiveness  of 
the  apparatus  depends  principally  upon  the  thoroughness 
with  which  this  detail  of  its  operation  is  fulfilled.  (3)  To 
purify  the  mixture  of  feed-water  and  condensed  exhaust  steam 


SEC.  257] 


FEED-WATER  HEATERS 


225 


by  filtration.  This  may  be  accomplished  (Fig.  235)  by 
causing  the  heated  water  to  percolate  through  chambers  filled 
with  filtering  material.  (4)  To  afford  storage  space  for  the 


fxferncr/  Shell 


Separating  Chamber 
Inlet  To  HeaHr- 


Fia.  236.— Oil-Separating  Element  Of  Moffat  Open  Exhauat-Steam  Feed- Water  Heater. 

| Drying  Coils-.      .'.;•:'.  •'.:-.';::.;:.-;     Absorption  Ice 
'"Machine 


Wetter  Softening 
Chemicals- — 

* 'Evaporator 


Exhaust  From  Engines,  , 
Pumps  Eft. ' 

From  Water  Main- * 


Boiler  Feed-Pump-1 
^•Feed-Water  Heater 
£  --Jo  Sewer      Boilers— 


Fio.  237.— Diagram  Showing  How  A  Feed- Water  Heater  Serves  As  A  Clearing  House 
For  All  Available  Supplies  Of  Exhaust  Steam  And  Water  Which  Are  Suitable  For 
Boiler  Feeding.  (Light  lines  represent  exhaust-steam  piping;  heavy  lines,  water 
piping.) 

heated  and  filtered  water  and  act  as  a  receiver  for  condensate 
from  various  sources  (Fig.  237). 

EXPLANATION. — The  feed-water  enters  the  heater  (Fig.  235)  through 
the  pipe  F.     The  rate  of  flow  is  controlled  by  the  valve  V,  which  is  oper- 
15 


226 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


ated  by  the  float  H.  The  water  rains  down  through  the  perforated 
plate  R  and  passes  successively  through  the  filter  beds  MI  and  M.  From 
the  filter  bed  M,  the  water  rains  down  through  chamber  A,  whence  it 
percolates  upward  through  the  coke  filter  in  chamber  N,  and  thence 
through  the  strainer  L  into  the  storage  chamber  Y. 


Co/d  WaHr 


Gravity 
Returns 


Fia.  238.— Cochrane  Open  Induction  Feed-Water  Heater. 

The  exhaust  steam  enters  the  heater  (Fig.  235)  through  the  nozzle  E, 
and  (Fig.  236)  is  diverted  to  a  downward  flow,  through  the  cups  C,  into 
the  separating  chambers  S.  The  momentum  of  the  oil-particles  precipi- 
tates them  to  the  bottoms  of  these  chambers.  As  the  oil  accumulates, 
it  flows  through  the  openings  (R,  Fig.  236)  into  the  space  surrounding 


SEC.  258]  FEED-WATER  HEATERS  227 

the  separating  chambers,  and  thence  out  through  the  drain-pipe  W.  The 
steam  (Fig.  235)  circulates  upward  through  the  core-pipe,  K,  and  is 
deflected  by  the  plate,  D,  through  lateral  openings  in  the  core-pipe,  into 
the  annular  chamber  A.  A  portion  of  the  steam  is  condensed  by  the 
water  percolating  through  the  filter  bed  M,  another  portion  ascends 
through  the  duct  T,  while  a  considerable  portion  reenters  the  core-pipe, 
K,  through  the  openings  above  the  deflecting  plate  D.  The  steam  which 
reenters  the  core-pipe  is  deflected  into  the  annular  chamber  A{  by  the 
plate  DI.  The  same  events  which  followed  the  entrance  of  the  steam 
into  chamber  A  then  ensue.  Some  of  the  steam  is  condensed  in  the  filter 
bed  Mi,  some  of  it  passes  up  through  the  duct  Ti,  while  the  remaining 
portion  again  reenters  the  core-pipe  through  the  openings  above  deflect- 
ing plate  DI.  The  steam  ascending  through  the  core-pipe  finally  en- 
counters the  cold  water  supply  as  it  trickles  down  through  the  rain-plate 
R.  Then  if  the  heater  is  operated  on  the  through  principle  the  uncondensed 
steam  passes  around  the  edge  of  the  upper  baffle,  U,  and  out  through 
the  nozzle  O.  With  induction  operation  the  exhaust  outlet,  0,  is  closed 
except  for  a  small  vent  pipe  leading  back  to  the  exhaust  pipe  (Fig.  227). 

The  perforated  pipes  B  and  X  have  external  connections,  through  the 
shell,  to  a  source  of  water  under  pressure.  Pipe  B  is  provided  for  flush- 
ing down  the  coke  filter.  Pipe  X  is  provided  for  washing  the  sludgy 
deposits  from  beneath  the  coke  filter  out  through  the  blow-off  valve. 

NOTE. — THE  CONDENSATE  FROM  A  GRAVITY  HEATING  SYSTEM  may 
be  piped,  G  (Fig.  235),  directly  to  an  open  feed-water  heater.  See  also 
Fig.  237. 

NOTE. — The  same  operations  are  performed  with  different  construction 
by  the  Cochrane  heater  (Fig.  238). 

258.  If  The  Carbonates  Of  Lime,  Magnesia  And  Iron  Are 
Dissolved  In  A  Feed  Water,  they  may  be  removed  by  an 
open    feed-water    heater.     These    impurities    precipitate    at 
temperatures  below  212  deg.  fahr.     Hence,  if  this  temperature 
is  maintained  in  an  open  feed-water  heater  the  impurities 
mentioned  will  be  deposited  in  the  heater.     Thus  the  forma- 
tion of  scale  in  the  boilers  may  be  largely  avoided.     For  the 
destructive  effects  of  scale  on  boiler  tubes  and  plates  see  the 
author's  STEAM  BOILERS. 

259.  Only  Liquid  OH  Can  Be  Removed ByThe  Oil-Separator 
Of  An  Open  Feed-Water  Heater. — Hence,  if  a  low  grade 
of  oil  is  used  for  engine-cylinder  lubrication,  the  separation 
may  not  be  complete.     This  will  be  due  to  the  fact  that  some 
of  the  constituents  of  low  grade  oils  vaporize  at  the  steam 
temperature.     The  oil  vapor  will  then  pass  into  the  heater 
and  form  an  emulsion  with  the  water.     Thus  a  portion  of  the 


228  STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 

oil  will  be  delivered  to  the  boiler.  Therefore,  none  but  a 
high  grade  of  oil  should  be  used  for  engine-cylinder  lubrication 
where  the  exhaust  from  the  engine  is  to  be  condensed  in  an 
open  feed-water  heater.  See  the  Author's  STEAM  BOILERS. 

NOTE. — OILY  FEED  WATER  Is  VERY  OBJECTIONABLE.  Oil  is  a  very 
poor  conductor  of  heat.  Hence,  if  the  oil,  which  may  be  admitted  to  a 
boiler  with  the  feedwater,  lodges  on  the  fire-sheets  or  tubes,  overheating 
of  the  sheets  or  tubes  may  result.  The  overheating  may  then  cause  the 
plates  or  tubes  to  bag  or  bulge,  thus  weakening  the  material  and  inviting 
rupture.  (See  the  author's  STEAM  BOILERS.)  Hence,  removal  of  the 
oil  from  the  exhaust  steam  which  is  used  is  a  very  important  function  of 
the  open  feedwater  heater. 

260.  The  Air  And  Carbonic  Acid  Gas  Which  Water  For 
Boiler-Feed  Generally  Holds  In  Solution  are  largely  liberated 
in  an  open  feed-water  heater  at  about  210  deg.  fahr.     If  the 
separation  takes  place  in  the  heater  no  damage  will    result. 
The  liberated  air  and  carbonic  acid  gas  will  pass  out  through 
the  vent  to  the  atmosphere.     But,  in  the  absence  of  an  open 
heater,  if  the  separation  takes  place  in  the  boiler,  the  liberated 
gases  will  combine  chemically  with  the  material  of  its  construc- 
tion and  rapid  corrosion  will  result. 

261.  The  Use  Of  A  Feed-Water  Heater  Is  Advisable  As  A 
Boiler  Protective  Measure  Even  Where  No  Economic  Saving 
Is  Apparent. — The  strains  in  boiler  plates,  due  to  cold  feed- 
water   striking   directly   against   them,    are   estimated    (The 
Locomotive)  at  8,000  to  10,000  Ib.  per  sq.  in.     This  in  addition 
to  the  normal  strain  produced  by  steam  pressure  is  quite 
enough  to  tax  the  girth  seams  beyond  their  elastic  limit  if  the 
feed  pipe  discharges  anywhere  near  them.     Hence,  it  is    not 
surprising  that  girth  seams  develop  leaks  and  cracks  in  99 
cases  out  of  every  100  in  which  the  feed  discharges  directly 
against  the  fire  sheets.     From  the  foregoing  it  is  evident   that 
the  feed-water  heater  is  a  necessary  part  of  the  equipment  of  a 
power  plant  aside  from  all  purely  economic   considerations. 

262.  The  Temperature  To  Which  Feed  Water  May  Be 
Raised   By   Steam  In  An  Open  Heater   depends   upon  the 
quantity  of  exhaust  steam  available,  the  initial  temperature 
of  the  feed-water,  and  the  temperature  of  the  exhaust  steam. 
When  all  of  the  exhaust  steam  which  is  delivered  to  the  heater 


SEC.  263]  FEED-WATER  HEATERS  229 

is  condensed  therein  the  final  temperature  of  the  feed  water 
may  be  computed  by  the  following  formula: 

r/iW/  +  0.9WS  (H  +  32) 
(77)    Tfz  =  W/  +  0  9W.  (degrees  Fahrenheit) 

Wherein  T/2  =  the  temperature  of  the  water  leaving  the 
heater,  in  degrees  Fahrenheit.  T/i  =  the  temperature  of 
the  water  entering  the  heater,  in  degrees  Fahrenheit.  W/ 
=  the  weight  of  the  feed-water  entering  heater,  in  pounds  per 
hour.  Ws  =  the  weight  of  the  exhaust  steam>  in  pounds  per 
hour.  H  =  the  total  heat,  above  32  deg.  fahr.  in  the  exhaust 
steam,  in  British  thermal  units  per  pound.  0.9  =  90  per  cent. 
=  the  assumed  efficiency  of  the  heater. 

NOTE. — When  the  result  obtained  by  For.  (77)  is  a  temperature 
greater  than  the  temperature  of  the  exhaust  steam,  it  means  that  all  of 
the  steam  will  not  be  condensed.  The  temperature  of  the  discharge 
from  the  heater  is,  then,  within  2  to  5  deg.  fahr.  of  the  exhaust  steam 
temperature,  and  the  amount  of  steam  condensed  may  be  calculated  by 
For.  (78). 

EXAMPLE. — A  1,200  h.p.  condensing  engine  uses  20  Ib.  of  steam  per  h. p. 
per  hr.  The  auxiliaries  use  2,400  Ib.  of  steam  per  hr.  The  exhaust  from 
the  auxiliaries  is  condensed  in  a  through-type  open  feed-water  heater. 
The  atmospheric  relief -valve  above  the  heater  is  set  for  a  back-pressure  of 
4  Ib.  per  sq.  in.  The  feed-water  is  delivered  from  the  hot-well  to  the 
heater  at  a  temperature  of  110  deg.  fahr.  What  is  the  temperature  of 
the  water  flowing  from  the  feed-water  heater  to  the  feed-pump? 

SOLUTION. — The  quantity  of  water  delivered  to  the  heater  =  (1,200  X 
20)  =  24,000  Ib.  per  hr.  As  given  in  a  table  of  the  properties  of  saturated 
steam,  the  total  heat,  above  32  deg.  fahr.,  in  steam  at  a  pressure  of  4  Ib. 
per  sq.  in.,  gage,  is  1,155  B.t.u.  per  Ib.  Hence,  by  For.  (77),  the  temper- 
ature of  the  water  leaving  the  heater  =  Tf2  =  [T/i  W/  +  0.9W.,(#  +  32)] / 
(W/  +  0.9  W.)  =  { (1 10 X  24,000)+  [0.9 X  2,400X  (1,155+32)]}  -=-  [2,4000 
+  (0.9  X  2,400)]  =  199  deg.  fahr. 

263.  In  A  Non-Condensing  Plant  Only  About  One-Seventh 
Or  Fourteen  Per  Cent.  Of  The  Steam  Exhausted  From  The 
Engine  And  Auxiliaries  Can  Be  Utilized  For  Feed-Water 
Heating;  About  Eighty-Six  Per  Cent.  Of  The  Exhaust  Steam  is 
Wasted. — The  feed  water  should  usually  be  heated  to  212  deg. 
fahr.  It  is  impossible  to  heat  it  to  a  higher  temperature  at 
atmospheric  pressure  without  causing  it  to  vaporize  into  steam. 
And,  furthermore  it  is  an  impossibility  to  heat  the  feed  water 


230  STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 

to  a  temperature  higher  than  that  of  the  exhaust  steam  which 
is  used  for  the  heating.  The  temperature  of  this  exhaust 
steam  is  always,  at  atmospheric  pressure,  212  deg.  fahr. 

NOTE. — THE  EXHAUST  FROM  THE  ENGINE  AND  AUXILIARIES  Is 
PRACTICALLY  ALL  STEAM,  although  it  carries  some  condensed  water. 
This  exhaust  steam  holds  the  same  amount  of  heat  as  any  steam  at  212 
deg.  fahr.  Now  the  latent  heat  in  this  steam,  the  heat  which  each  pound 
of  steam  will  give  up  in  changing  from  steam  at  212  deg.  to  water  at 
212  deg.  is,  as  taken  from  a  steam  table,  970.4  B.t.u.  But  the  heat 
required  to  raise  the  temperature  of  1  Ib.  of  water  from  50  deg.  fahr. 
(which  is  the  average  cold  feed-water  temperature)  to  212  deg.  fahr.  is 
only:  212  —  50  =  162  B.t.u.  Therefore,  the  number  of  pounds  of  cold 
feed  water  which  will  be  heated  from  50  deg.  fahr.  to  212  deg.  fahr.  by 
1  Ib.  of  exhaust  steam  will  be  970.4  +  162  =  6  Ib.  One  Ib.  of  steam  will, 
then,  afford  all  of  the  heat  that  6  Ib.  of  feed  water  can,  under  the  circum- 
stances, absorb. 

264.  The  Proportion  Of  The  Total  Steam  Generated  In  A 
Non-Condensing  Plant  Which  Is  Useful  In  Feed  Water  Heat- 
ing is  about  14  per  cent.     For  each  6  Ib.  of  cold  water  at  50 
deg.  fahr.  (as  above  described)  which  is  pumped  into  the  boiler 
1  Ib.  of  water  condensed  from  exhaust  steam  is  pumped  in  with 
it.     (This  assumes  that  an  open  feed-water  heater  is  used). 
This  gives  a  total  of  7  Ib.  of  hot  feed  water  pumped  into  the 
boiler  for  each  pound  of  exhaust  steam  used.     Thus  (See  also 
Sec.  263)  only  about  K  or  14  percent,  of  the  total  water  pumped 
into  the  boiler  (that  is,  %  of  the  steam  generated  but  finally 
exhausted  through  the  engine  and  auxiliaries)  can  be  effective 
for  feed-water  heating.     The  remainder,  or  86  per  rent,  of  the 
exhaust  steam  is  wasted — unless  it  is  employed  for  room  heat- 
ing or  some  similar  useful  non-power-generation  purpose. 

265.  In  A  Condensing  Plant  Which   Carries    A   26-Inch 
Vacuum  Only  About  One-Eleventh  Or  Nine  Per  Cent.  Of 
The  Steam  Generated  By  The  Boiler  Can  Be  Used  For  Heat- 
ing The  Feed  Water. — In  a  condensing  plant  all  of  the  steam 
from  the  engine  is  condensed  with  cold  water  and  is  discharged 
into  the  hot  well.     Some  of  the  auxiliaries  should  be  operated 
non-condensing  so  that  their  exhaust  can  be  used  for  heating 
the  feed  water  from  the  hot  well  up  to  212  deg.  fahr.  if  possible. 
The  temperature  of  this  condenser-discharge  water  which  is 
thus  used  from  the  hot  well  for  boiler  feed  is  (with  a  26-in. 


SEC.  266]  FEED-WATER  HEATERS  231 

vacuum)  about  120  deg.  fahr.  Therefore,  to  raise  its  temper- 
ature to  212  deg.  fahr.  there  will  be  required  only  212-120  = 
92  B.t.u.  It  is  assumed  that  an  open  feed-water  heater 
will  be  used.  Hence,  for  these  conditions  the  number  of 
pounds  of  feed  water  which  will  be  heated  from  120  deg.  to  212 
deg.  by  1  Ib.  of  exhaust  steam  (which  will  give  up  970.4  B.t.u. 
of  latent  heat  in  changing  from  steam  at  212  deg.  to  water  at 
212  deg.)  will  be:— 970.4  -r-  92  =  10.6  Ib.  That  is,  1  Ib.  of  the 
exhaust  steam  at  212  deg.  fahr.  will  heat  10.6  Ib.  of  the  120 
deg.  fahr.  feed  water  to  212  deg.  fahr.  How,  with  each  pound 
of  the  hot- well  water  which  is  fed  into  the  boiler,  the  1  Ib.  of 
condensed  steam  which  is  used  in  raising  the  temperature  of 
the  hot- well  water  is  fed  in  with  it.  Hence,  for  each  1  Ib.  of 
exhaust  steam  utilized  for  feed-water  heating  there  is  fed  into 
the  boiler: — 10.6  +  1  =  11.6  Ib.  of  feed  water  at  a  temperature 
of  212  deg.  F. 

This  being  true,  there  is  only  1/11.6  =  8.6  per  cent,  or  say, 
9  per  cent,  of  the  total  steam  generated  by  the  boiler  which  can 
be  used  for  heating  feed  water.  Obviously,  then,  the  ideal 
economic  condition  for  a  condensing  plant  which  carries  a  26 
in.  vacuum  is  to  have  auxiliaries  which  will  furnish  exhaust 
steam  to  an  amount  equivalent  to  about  9  per  cent,  of  the 
steam  generated  by  the  boiler.  It  should  be  understood  that 
the  9  per  cent,  is  the  ideal  value  which  applies  only  for  the 
water  temperature  conditions  specified  for  this  example. 
Losses  such  as  condensation  and  the  like,  for  which  no  allow- 
ance has  been  made  in  this  problem,  will  tend  to  increase  above 
9  per  cent,  the  amount  of  exhaust  steam  which  can  be  used  for 
feed-water  heating. 

NOTE. — IT  Is  REASONABLE  To  EXPECT  THAT  THE  AUXILIARIES  IN 
THE  AVERAGE  PLANT  WILL  SUPPLY  ABOUT  THE  AMOUNT  OP  EXHAUST 
STEAM  REQUIRED  FOR  HEATING  THE  FEED  WATER.  Every  effort  should 
be  exerted  to  produce  just  enough  exhaust  steam  to  heat  the  feed  water 
up  to  210  deg.  or  212  deg.  But  there  should  be  no  exhaust  in  excess  of 
this.  If  there  is  excess  exhaust  the  heat  in  it  will  be  wasted. 

266.  To  Compute  The  Weight  Of  Steam  Condensed  By  An 
Open  Heater,  use  the  following  formula: 


232 


STEAM  POWER  PLANT  AUXILIARIES 


[Drv.  7 


Wherein:  Ws  =  weight  of  steam  condensed,  in  pounds  per 
hour.  T/2  =  discharge  temperature  of  feed-water  in  deg.  fahr. 
Tfi  =  initial  temperature  of  feed-water  in  deg.  fahr.  WF  = 
weight  of  hot  water  delivered  by  heater  in  pounds  per  hour. 
H  =  the  total  heat  in  the  exhaust  steam,  above  32  deg. 
fahr.,  in  British  thermal  units  per  pound.  0.9  =  90  per  cent, 
which  is  the  assumed  efficiency  of  the  heater. 

EXAMPLE. — Suppose  2,400  Ib.  per  hr.  of  feed  water  is  required  by  a 
boiler.  Steam  at  227  deg.  fahr.  is  available  for  feed-water  heating. 
The  initial  temperature  of  the  feed-water  is  90  deg.  fahr.  and  it  is  delivered 
at  212  deg.  fahr.  What  weight  of  steam  is  condensed  by  the  heater? 

SOLUTION. — As  given  in  a  table  of  the  properties  of  saturated  steam, 
the  total  heat  above  32  deg.  fahr.,  in  steam  at  227  deg.  fahr.  is  1,156 
B.t.u.  per  Ib.  By  For.  (78)  the  weight  of  steam  condensed, 


=  293  Ib.  per  hr. 


I  -T  n  +  0.1T/1 

(212  -  90)2,400 
0.9(1,156  +  32)  -  90  +  (0.1  X  212) 

.--Crane     .-Wafer  frr/ef$—~*  Live  Steam. 
Connection'. 


FIG.  239.— Hoppes  Live-Steam  Heater  And  Purifier  With  Head  Removed  And 
Hanging  On  The  Crane.  (Heaters  of  very  similar  design  are  used  for  regular  ex- 
haust-steam heating  service.) 

267.  The  Pan  Or  Tray  Area  Required  In  An  Open  Heater 

using  pans  or  trays  (Fig.  239  and  240)  is  (KENT'S  MECHAN- 
ICAL ENGINEERS'  POCKETBOOK)  as  follows: 


Quality  of  water 

Surface  in  sq.  ft.  pe 
heated  p 

r  1,000  Ib.  of  water 
er  hour 

For  vertical  type 

For  horizontal  type 

Very  bad  water  
Medium  muddy  water  

8.5 
6  0 

9.1 

6  5 

Clear  water  little  scale  

2.0 

2.2 

SEC.  267] 


FEED-WATER  HEATERS 


233 


NOTE. — The  practice  in  heater  manufacture  is,  however,  to  use  a  total 
tray  surface  equal  to  about  3  to  4  times  the  horizontal  sectional  area  of 
the  shell  at  the  plane  at  which  the  trays  are  located.  The  space  between 
the  pans  or  trays  is  made  not  less  than  0.1  the  width  for  rectangular  and 
0.25  times  the  diameter  for  round,  trays  or  pans.  It  is  not  customary  to 


Spray  Holes--.,, ,  Sepotrcrt/ngf  Baffles-, 


Exhorusf-Stecrm  Ouf/ef-^     frays- 
Door  For  Removing          . 


Exhctust-\ 
5teo/m  • 
Inlet-' 


Oil 


Pump'- 
5ucf/6n 


Fia.  240. — Blake-Knowles   Open  Exhaust-Steam  Feed-Water   Heater   Using   Inclined 

Trays. 

use  more  than  six  pans  in  a  tier.  The  size  of  the  water  storage  or  settling 
space  in  the  horizontal  type  varies  from  0.25  to  0.4  the  volume  of  the 
shell;  and  in  the  vertical  type  from  0.4  to  0.6.  The  niters  occupy  from 
10  to  15  per  cent,  of  the  volume  of  the  shell  in  the  horizontal  type  and 
from  15  to  20  per  cent,  in  the  vertical. 


234  STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 

268.  To  Compute  The  Approximate  Size  Of  Shell  Required 
For  An  Open  Heater,  use  the  following  formulae: 

(79)  Af=^f  (square  feet) 

A-L/t 
or 

(80)  L"=Jjc  (feet) 

Wherein:  A/  =  cross-sectional  area  of  heater,  in  square  feet. 
Lh  =  height  of  heater,  in  feet.  W/  =  weight  of  feed-water 
heated  by  the  heater,  in  pounds  per  hour.  K  =  a  constant; 
for  clear  water  K  =  270;  for  slightly  muddy  water  K  =  200; 
and  for  very  muddy  water  K  =  70.  The  formula  is  based  on 
proportions  of  commercial  heaters  furnishing  6,000  or  more 
Ibs.  per  hour  of  feed-water.  These  heaters  were  all  of  upright 
design  having  Lh  not  more  than  3  times  the  smaller  base 
dimension.  For  heaters  furnishing  3,000  to  5,000  Ib.  per  hr., 
allow  25  per  cent,  more  capacity  than  given  by  the  formula. 

EXAMPLE. — What  should  be  the  tray  area  and  shell  dimensions  of  an 
open  heater  to  heat  10,000  Ib.  per  hr.  of  feed  water.  The  heater  is  to  be 
square  in  cross  section  and  the  height  is  to  be  twice  the  base  dimension. 
The  water  is  slightly  muddy. 

SOLUTION. — By  For.  (80),  Lh  =  • — J-   =  -  -  or  AfLh=  50cu.fl. 

A.  /*»-         •**•  j  '^  <wv/iJ 

But,  for  a  square  section  one  side  being  %Lh,  %Lh  X  ^Lh  X  Lh  =  50  or 
Lh  =  \/4  X  50  =  5.85  ft.  or  about  5  ft.,  10  in.  The  base  is  2  ft.  11  in. 
or  2.92  ft.  square.  The  tray  area  required  is  (Sec.  267)  about  3.5  times 
the  cross-sectional  area,  or  2.92  X  2.92  X  3.5  =30  sq.  ft.  approx. 
Assuming  that  six  trays  are  to  be  used  they  will  beabout  v730/6  or  2ft. 
3  in.  square.  They  should  be  at  least  0.1  the  width  or  2%  in.  apart. 


SEC.  2691 


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FEED-WATER  HEATERS 

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236 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 


270.  Table  Of  General  Data  And  Approximate  Net  Selling 
With  Exhaust   Steam   (Harding  and   Willard,   MECHANICAL 


Horsepower  rating 

50 

100 

150 

200 

250 

300 

350 

425 

500 

Pounds  feed  water  per  hour  
Weight  in  pounds  
Net  price  f.o.b.  factory  

1500 
1200 
$102 
25 
21 
62 
4 
1 

IK 
in 

4 
17 
12 

3000 
1300 
129 
27 
23 
63 
5 
1 
2 
IK 
4 
19 
13K 

4500 
1800 
159 
30 
25 
70 
6 
IK 
2K 
2 
4 
21 
15 

6000 
2100 
188 
32 
27 
73 
6 
IK 
2K 

2 
5 
22 
15 

7500 
2400 
229 
34 
29 
78 
7 
IK 
3 
2K 
5 
24 
16K 

9000 
2700 
256 
43 
29 
78 
7 
IK 
3 
2K 
5 
24 

16K 

10500 
3000 
275 
39 
33 
84 
8 
2 
4 
3 
5 
28 
21 

12750 
3300 
302 
49 
33 
84 
8 
2 
4 
3 
5 
28 
21 

15000 
3700 
331 
45 
38 
75 
9 
2 
4 
3 
10 
32 
10K 

TV  idth   inches 

Depth   inches 

Height,  inches  

Max.  dia.  exh.  inlet  and  outlet  
Dia.  cold  water  supply 

Dia.  ins   pump  suction 

Dia.  waste  and  overflow 

Number  of  trays         .                         .  . 

Length  per  tray,  inches  

Width  per  tray,  inches.  .  . 

NOTE. — The  heaters  tabulated  above  are  designed  for  power-plant  operation,  and  not 
See  notes  below  Table  271  regarding  prices.  For  estimating  purposes  and  preliminary 
10  per  cent  to  cover  steam  consumption  of  auxiliaries  (pumps,  etc.).  The  value  so 
hour."  Select  a  heater  accordingly.  In  considering  heaters  of  the  same  general  type, 

271.  Table  Of  General  Data  And  Approximate  Net  Selling 
With  Exhaust  Steam.  (Harding  and  Willard,  MECHANICAL 


Horsepower  rating 

50 

60 

70 

80 

100 

130 

160 

200 

2iO 

Pounds  feed  water  per  hour 

1500 

1800 

2100 

2400 

3000 

3900 

4800 

fiOOO 

7200 

Tube  heating  surface,  »o    ft 

17 

20 

23 

27 

33 

43      53      f»7 

80 

Number  of  tubes 

18 

18 

18 

18 

18 

36 

36 

36 

36 

Diameter  of  tubes,  inches  

IK 

IK 

IK 

IK 

IK 

IK 

IK 

IK 

IK 

Length  of  tubes,  inch 

PR     .  . 

35% 

42>£ 

49^ 

56  K 

70 

45^ 

56 

69  % 

83  K 

Diameter  of  shell,  inches  

12 

12 

12 

12 

12 

16 

16 

16 

16 

Diameter  of  feed  pipe,  inches  

IK 

IK 

IK 

IK 

IK 

2 

2 

2 

2 

Diameter  of  exhaust  pipe,  inches  

6 

6 

6 

6 

6 

8 

8 

8 

8 

Total  length  —  horizontal  heater  

4'  7" 

5'  2" 

5'  8" 

6'  3" 

7'  5* 

5'  7* 

6'  5" 

7'  7' 

8'  8" 

Total  length 

vertical  type  .  .  . 

5'  4" 

5'  11* 

6'  6" 

7'0" 

8'  2" 

6'  4* 

7'  2" 

8'  4" 

9'  5" 

on  legs 

horizontal  type 

2'  6" 

2'  6" 

2'  6"  2'  6* 

2'  6" 

3'0" 

3'0" 

3'0" 

3',0* 

Shipping  weight, 

vertical  type  .  .  . 

880 

900 

950 

1000 

1125 

1250 

1550 

1700 

1900 

pounds 

horizontal  type 

950 

1000 

1050 

1250 

1400 

1675 

1750 

1900 

2000 

Net  selling 

vertical  type.  .  . 

$133 

140 

147 

154 

168 

193 

214 

235 

252 

price 

horizontal  type 

$144 

155 

163 

171 

186 

214 

238 

260 

280 

NOTE. — "Closed"  feed  water  heaters  are  either  of  the  water-tube  or  steam-tube  type: 
exhaust  steam  passing  through  the  shell.  In  the  latter  the  exhaust  steam  is  passed 
shell.  The  water-tube  heater  is  the  type  generally  used  in  steam-power-plant  work. 
Heaters  may  be  vertical  or  horizontal  type  as  space  dictates.  See  note  under  Table 
note  (a)  material  of  tubes;  (b)  square  feet  of  tube  heating  surface;  (c)  the  weights;  (d)  the 

NOTE. — The  prices  listed  above  are  for  1916  and  cannot  be  relied  upon  closely  at 


SEC.  271] 


FEED-WATER  HEATERS 


237 


Prices  Of  Feed -Water  Heaters  Of  The  Open  Type— For  Use 

EQUIPMENT  OF  BUILDINGS,  Vol.  II) 


600 

750 

850 

1000 

1250 

1500 

1750 

2000 

2500 

3000 

4000 

5000 

6000 

18000 

22500 

25500 

30000 

37500 

45000 

52500 

60000 

75000 

90000 

120000 

150000 

180000 

4300 

4900 

5400 

6400 

7000 

8300 

9100 

10000 

11000 

12000 

15000 

16000 

17000 

380 

420 

493 

540 

618 

720 

820 

925 

1060 

1155 

1410 

1605 

1738 

55 

50 

60 

56 

68 

67 

78 

113 

113 

115 

128 

130 

132 

38 

42 

42 

48 

47 

56 

53 

42 

48 

54 

54 

62 

70 

75 

84 

84 

87 

84 

97 

97 

88 

88 

88 

100 

100 

100 

10 

10 

12 

12 

12 

14 

14 

16 

16 

18 

20 

22 

24 

2K 

2K 

2K 

2H 

3 

3 

3 

3K 

3K 

4 

4K 

5 

5H 

4 

4 

5 

5 

5 

5 

6 

6 

7 

8 

9 

10 

10 

3K 

3K 

3K 

3K 

4 

4 

4 

5 

5 

6 

7 

8 

8 

10 

10 

10 

20 

20 

20 

20 

20 

40 

40 

40 

40 

40 

32 

36 

36 

22 

22 

25 

25 

36 

22 

25 

25 

29 

33 

IOK 

12 

12 

15 

15 

18 

18 

15 

15 

15 

18 

18 

18 

designed  to  operate  in  conjunction  with  steam-heating  systems  under  back  pressure, 
determinations,  compute  the  steam  consumption,  per  hour  of  the  main  engines,  and  add 
obtained  corresponds  to  the  line  of  the  table  entitled  "pounds  of  feed  water  heated  per 
but  of  different  manufacture,  compare  particularly  cubic  contents,  weights,  and  prices. 

Prkes  Of  Feed -Water  Heaters  Of  The  Closed  Type— For  Use 

EQUIPMENT  OF  BUILDINGS,  VOL.  II) 


300 

350 

400 

500 

600 

700 

800 

900 

1000 

1200 

1500 

1800 

2000 

9000 

10500 

12000 

15000 

18000 

21000 

24000 

27000 

30000 

36000 

45000 

54000 

60000 

100 

117 

133 

167 

200 

233 

266 

300 

333 

400 

500 

600 

667 

60 

60 

60 

90 

90 

90 

126 

126 

126 

126 

150 

150 

186 

IK 

IK 

iH 

IK 

IK 

IK 

IK 

IK 

IK 

IK 

1% 

1% 

IK 

62% 

73 

83K 

69^ 

83% 

96% 

78K 

88% 

97% 

117K 

112% 

135 

111% 

21 

21 

21 

25 

25 

25 

29 

29 

29 

29 

34 

34 

39 

2K 

*H 

2K 

3 

3 

3 

4 

4 

4 

4 

5 

5 

6 

10 

10 

10 

12 

12 

12 

16 

16 

16 

16 

18 

18 

22 

7'  3" 

8'1" 

9'0" 

8'  2" 

9'  4" 

10'  5" 

9'  2" 

10'  1" 

10'  10" 

12'  5" 

13'  5" 

15'  3" 

13'  10" 

8'  9* 

9'  7" 

10'  5" 

9'  7" 

10'  8" 

11'  10" 

10'  8" 

11'  7" 

12'  4" 

14'  0" 

14'  4" 

14'  2" 

13'  8" 

3'  5" 

3'  5" 

3'  5" 

3'  10" 

3'  10" 

3'  10" 

4'  3" 

4'  3" 

4'  3" 

4'  3" 

4'  11" 

4'  11" 

5'  6" 

2500 

2800 

2900 

3800 

4000 

4400 

5000 

5500 

5800 

6300 

7200 

11000 

14000 

2600 

2900 

3200 

4100 

4300 

4700 

5500 

6000 

6500 

7200 

10000 

12000 

14000 

322 

350 

378 

490 

540 

575 

660 

708 

750 

840 

1190 

1300 

1430 

356 

390 

420 

545 

598 

637 

730 

785 

830 

938 

1320 

1420 

1610 

In  the  former  the  feed  water  circulates  through  the  tubes  and  is  surrounded  by  the 
through  the  tubes  and  the  feed  water  (surrounding  th«  tubes)  is  carried  through  the 
The  shell  is  usually  of  cast  iron  and  brass  or  copper  tubes  are  almost  always  used. 
270  regarding  prices,  and  estimation  and  selection  of  heater.  In  making  comparisons, 
prices. 
present. 


238 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


272.  General  Rules  For  Selecting  Exhaust  Steam  Feed- 
Water  Heaters  are:  Use  an  open  heater  whenever  possible 
on  account  of  its  greater  efficiency  as  a  heater  and  purifier 
and   ease    of    cleaning.     It    cannot    be   used:  (1)  When    the 
feed-water  in  the  heater  must  be  under  a  pressure  of  more  than 

about  5  Ib.  persq.  in.  (2)  When 
the  steam  used  for  heating  is  ex- 
hausted under  a  vacuum  as  in 
condensing  operation.  (3) 
When  the  feed-water  must  be  kept 
entirely  free  of  oil.  (4)  When  the 
feed-water  heater  is  connected  to 
the  feed  pump  between  the  pump 
and  the  boilers.  Under  any  of 
the  four  conditions  listed  a  closed 
heater  must  be  used. 

NOTE. — THE  EFFECTIVENESS  OF  AN 
OPEN  FEED-WATER  HEATER  As  A 
PURIFIEB  depends  not  alone  upon  the 
area  of  heating  surface  which  it  con- 
tains, but  also  upon  its  volume  of 
water-storage  capacity.  Storage  ca- 
pacity is  variable  to  a  greater  extent 
than  is  heating  surface.  If  the  water 
is  hard,  purification  is  desirable.  The 
longer  the  water  remains  in  the  heater, 
the  more  thorough  will  be  the  pre- 
cipitation. Hence,  a  larger  water- 
storage  space  is  required  than  would 
otherwise  be  necessary.  On  the  other 

hand,  the  heater  may  be  used  in  a  surface-condensing  plant,  where 
the  condensate,  which  is  usually  free  from  scale-forming  impurities  is 
used  as  feed  water.  Then,  if  there  is  a  fairly  uniform  load,  the  con- 
densate is  delivered  to  the  heater  at  a  uniform  rate,  and  only  such 
volume  of  water  need  be  carried  in  storage  as  will  insure  a  steady  supply 
to  the  feed-pump. 

273.  Closed  Exhaust-Steam  Feed-Water  Heaters  May  Be 
Grouped  Into  Two  Classes :  (1)  Water-tube  heaters  (Fig.  241) 
in  which  the  feed-water  passes  through  a  set  of  brass  or  copper 
tubes   which   are    surrounded    by   the    exhaust    steam.     (2) 
Steam-tube  heaters   (Fig.   242)   in  which  the  exhaust  steam 


FIG.    241. — The    National    Coil   Type 
Closed  Feed- Water  Heater. 


SEC.  273] 


FEED-WATER  HEATERS 


239 


Hot-Muter 
Outlet-. 


-IT? 


'-Dram 


FIG.  243.— Diagram  Of  Parallel-Cur- 
rent Return-Flow  Closed  Feed-Water 
Heater. 


Steam 


'  Cohoterisation Drip-Pipe  i 

To  Sfeo/m  Space 

FIG.  242.— Steam-Tube  Type  Of  Closed          FIG.  244.— Diagram  Of  Counter-Cur- 
Exhaust-Steam  Feed- Water  Heater.          rent    Return-Flow    Closed    Feed-Water 

Heater. 


l-Sec  t  i  o  n  CO 
(Water  How) 


E-S0ctionB 
(Steam  Flow) 


HE- Sect  ion    AA 
(Water  Flow) 


FIG.  245. — Cross  Sections  Of  Blake-Knowles  Heater   (Fig.  220)  Showing  Multi-Flow 

Arrangement. 


240 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


passes  through  a  set  of  brass  or  copper  tubes  which  are  sur- 
rounded by  the  feed  water. 

NOTE. — Closed  feed-water  heaters  may  be  designed  so  that  (Fig.  243) 
the  water  and  steam  flow  in  the  same  direction,  or  (Fig.  244)  in  opposite 
directions.  In  the  first  case,  the  heater  is  called  a  parallel-current  heater. 
In  the  second  case  it  is  called  a 
counter-current  heater.  If  the 
heater  is  so  built  that  the  water 
flows  straight  through,  it  is  called 
a  single-flow  heater.  If  the  water 
flows  back  and  forth  through  the 
tubes  a  number  of  times  (Figs.  220 
and  245)  it  is  called  a  multi-flow 
heater.  If  the  water  flows  through 
coiled  tubes  (Fig.  241)  it  is  called 
a  coil  heater.  If  the  water  is  forced 
across  the  heating  surface  in  a  thin 
sheet  or  film  it  is  called  a  film 
heater. 

274.  The  Tubes  In  A  Closed 
Feed-Water  Heater  May  Be 
Either  Straight  (Fig.  220)  Or 
Spirally  Corrugated  (Fig. 
246).  It  is  claimed  for  the 
corrugated  construction  that 
the  spiral  flow  of  the  water 

thn  Of  Cast-Iron  Head 


'-Drain  Openi'rxf 


FIG.  247. — Schutte  and  Koerting  Ver- 
tical Straight-Tube  Closed  Heater — Multi- 
Flow  Type. 


ferrule,  For  Expanding 
Tube-End 


FIG.  246. —  End  Of  Copper  Corru- 
gated Tube  In  Wainwright  Closed  Feed- 
Water  Heater. 


through  the  tubes  increases  the  contact  pressure  between 
the  water  and  the  tube  surface,  thereby  facilitating  the  heat- 
transmission.  It  is  also  claimed  that  the  spiral  currents  of 
water  tend  to  scour  the  surfaces  and  prevent  the  accumulation 
of  scale  thereon. 


SEC.  275] 


FEED-WATER  HEATERS 


241 


275.  A  Corrugated  Heater-Tube  Gives  Greater  Heating 
Surface,  for  a  given  water  volume,  than  does  an  uncorru- 
gated  tube.  Further  advantages  claimed  for  corrugated 
tubes  (Fig.  246)  are:  they  give  a  higher  rate  of  conduction 
per  unit  length  than  smooth  tubes,  the  corrugations  take 
up  all  heat  strains  making  more  rigid  construction  of  the 
heater  possible.  Corrugated  tubes,  it  is  claimed,  are  prefer- 
able where  the  range  of  temperature  of  the  water,  between 
inlet  and  outlet,  is  extreme,  or  where  the  velocity  of  the  water 
through  the  heater  is  very  high. 

NOTE. — WHEN  STRAIGHT  UNCORRUGATED  TUBES  ARE  USED  IN  CLOSED 
HEATERS,  a  floating  head  arrangement  (H,  Fig.  247)  is  usually  used  to  allow 


.-Double  Coil  Of  Copper 
Tubing      **^ 


Mafkoib/e 
Iron-.. 

Clamp-; 


FIG.      248.  — Inside     Manifold     Of     Whitlock 
Double-Coil  Closed  Heater. 


FIG.  249. — How  The  Coils  Are 
Secured  In  A  National  Coil 
Heater. 


for  expansion  in  the  -tubes.  Where  the  tubes  are  bent  (Fig.  242)  or 
coiled  (Fig.  241)  this  feature  is  unnecessary  as  the  tubes  are  free  to 
expand  and  contract  without  straining  the  supporting  head.  Methods 
of  connecting  and  supporting  coiled  tubes  are  shown  in  Figs.  248  and  249. 

276.  Steam-Tube  Closed  Feed-Water  Heaters  (Fig.  242) 
are  designed  for  service  where  a  varying  demand  for  steam 
necessitates  very  irregular  feeding  of  the  boilers.  This  condi- 
tion might  exist  where  the  steam-using  apparatus  which 
supply  the  exhaust  steam  for  heating  the  water  are  operated 
intermittently.  By  sending  the  steam,  instead  of  the  water, 
through  the  tubes,  the  space  surrounding  the  tubes  is  available 
for  storage  of  a  comparatively  large  volume  of  heated  water 

16 


242 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


during  the  intervals  when  the  feed-valve  is  closed.  The 
water  stored  in  the  heater  may  then  absorb  the  heat  from 
the  intermittent  deliveries  of  exhaust  steam. 

NOTE. — AN  INTERMITTENT  DELIVERY  OF  EXHAUST  STEAM  To  THE 
FEED- WATER  HEATER  might  occur  in  a  plant  where  hydraulic-elevator 
pumps,  or  the  engines  for  operating  trip-hammers,  cotton  compresses, 
or  in  similar  irregular  service,  are  depended  upon  for  supplying  the  steam. 

277.  To  Compute  The  Tube  Heating-Surface  Required  For 
A  Closed  Exhaust-Steam  Feed-Water  Heater,  use  the  follow- 
ing formula: 


(81) 


U(T,.  - 


(square  feet) 


1000 


Wherein  A/  =  the  total  heating  surface  of  the  tubes,  in  square 

feet.  W/  =  the  weight  of  feed 
water  to  be  heated,  in  pounds 
per  hour.  T/i  =  the  tempera- 
ture of  the  water  entering  the 
heater,  in  degrees  Fahrenheit. 
Tf2  =  the  temperature  of  the 
water  leaving  the  heater,  in  de- 
grees Fahrenheit.  T '/,  =  the 
temperature  of  the  exhaust 
steam  in  degrees  Fahrenheit. 
U  =  the  coefficient  of  heat- 
transfer  for  the  surface,  in 
British  thermal  units  per  hour 
per  square  foot  per  degree  tem- 
perature difference  as  given  in 


Mia 


Water  Va!ocity-Ft.  Par 
FIG.  250. — Graph  Showing  Effect  Of 
Water  Velocity  On  Coefficient  Of  Heat 
Transfer    Through    Tubes    Of    Closed      Table  278. 
Feed- Water  Heaters. 

NOTE. — THE  COEFFICIENT  OF  HEAT  TRANSFER  IN  CLOSED  HEATERS 
VARIES  WITHIN  WIDE  LIMITS.  It  depends  mainly  upon  the  thickness 
and  composition  of  the  conducting  wall,  the  disposition  of  the  heating- 
surface,  the  water  velocity  through  the  heater  (Fig.  250)  and  upon  the 
conditions  under  which  the  heater  is  operated.  It  may  range  from  150  to 
1000  or  more.  The  first  of  these  values  may  be  obtained  with  a  steel- 
tube  heater  in  which  the  water-velocity  is  low.  The  second  may  be 
realized  with  corrugated  brass-tube  heaters,  of  the  film  type,  in  which 


SEC.  278]  FEED-WATER  HEATERS  243 

the  water-velocity  is  very  high.  The  values  given  in  Table  278  are  for 
commercial  designs 

EXAMPLE. — A  closed  exhaust-steam  feed-water  heater  is  required  to 
heat  10,000  Ib.  of  feed-water  per  hr.  from  60  to  196  deg.  fahr.  with  steam 
at  212  deg.  fahr.  The  heater  is  to  be  of  the  multi-flow  corrugated  brass- 
tube  type.  What  should  be  the  area  of  the  tubing? 

SOLUTION: — By  Table  278  U  =  400.  Hence,  by  For.  (81)  A/  =  W/ 
(Tf,  -  Tfl)/(U{Tft  -  M(Tfl  +  r,j})  =  10,000  X  (196  -  60)/ 
(400(212  -  ^[60  +  196]})  =  40.5  sq.ft. 

NOTE. — Increasing  the  velocity  of  the  water  passing  through  a  heater 
increases  (Fig.  250)  the  coefficient  of  heat  transmission.  In  order  to 
realize  the  possible  maximum  feed-water  temperature,  and  at  the  same 
time  use  a  moderately  high  velocity  of  flow,  the  tubes  should  be  as 
long  as  is  feasible,  and  of  small  diameter. 

278.  Table  Showing  Average  Coefficients  Of  Heat  Trans- 
mission In  Closed  Feed-Water  Heaters  (Gebhardt). 


Type  of  heater 


Average  coefficient 

of  heat-transfer  = 

U,  For.  (81) 


Single-flow,  steel  water-tube 

Single-flow,  plain  brass  water-tube 

Single-flow,  corrugated  brass  water-tube 

Spiral  coil,  plain  brass  water-tube 

Multi-flow,  plain  brass  water-tube 

Multi-flow,  corrugated  brass  water-tube 

Multi-flow,  plain  brass  water-tube,  with  retarders 

Film,  corrugated  water-tubes 

Multi-flow,  iron  steam-tube 

Multi-flow,  brass  steam-tube 

Multi-flow,  copper  steam-tube 


150 
200 
300 

350  to  700 
350 
400 
450 
600 

100  to  225 
200  to  450 
220  to  475 


279.  Closed  Exhaust-Steam  Feed -Water  Heaters  Are 
Sometimes  Rated  In  Terms  Of  Heater  Horsepower. — By 
using  For.  (81)  it  can  be  shown  that  one  square  foot  of  heater 
surface  will  suffice  to  heat  103.5  Ib.  of  water  per  hr.  from  60 
deg.  fahr.  to  194  deg.  fahr.,  with  a  coefficient  of  heat  transfer 
(Sec.  277)  of  about  165  B.t.u.  per  sq.  ft.  per  hour  per  deg. 
difference  in  temperature.  On  the  above  outlined  basis  and 
on  the  assumption  that  34.5  Ib.  of  feed  water  is  required  per 
boiler  horse-power  per  hour,  a  closed  heater  will  supply 
103.5  -s-  34.5  =  3  boiler  horse  power  per  sq.  ft.  of  heating 


244 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 


surface.     Hence:  1  -7-  3  =  >£    sq.    ft.    of    heater    surface    is 
sometimes  allowed  per  boiler  horsepower. 

NOTE. — DOUBLE  HEATER  INSTALLATIONS  (Fig.  251)  are  used  in  large 
power  plants  which  are  operated  continuously.  These  consist  of  two 
separate  feed-water  heaters  which  are  so  connected  as  to  receive  exhaust 
steam  from  a  common  exhaust  pipe  and  water  from  a  common  water- 
supply  pipe.  With  an  installation  of  this  kind,  one  heater  may  be  cut 


Exhaust  Out  let  To  Atmosphere— 
Oil  Separator^ 


.-Cut-Out  Votlre 


Cold  Wafer  Regu/at/ngr  Yor/ye 

I    -    P        I         a         n 


Exhaust  Outlet  To  Atmosphere-.. 
Water  Inlet- 

m 


•Water  Inlet 


#^^ 

^Feed-Pump  Suction  Ma/'n  Exhaust  Inlet-'  ^Overflow 

H -Elevation" 

FIG.  251. — A  Double  Heater  Installation. 

out  of  service  for  cleaning  or  inspection  while  the  other  continues  to 
supply  hot  water  for  the  boilers. 

280.  The  Relative  Advantages  And  Disadvantages  Of 
Open  And  Closed  Feed-Water  Heater  may  (See  Gebhardt's 
STEAM  POWER  PLANT  ENGINEERING)  be  summed  up  as  follows: 
(1)  With  an  open  heater  the  water  may  be  heated  to  the  temperature 
of  the  exhaust  steam.  With  a  closed  heater  the  possible  maximum 


SEC  281.]  FEED-WATER  HEATERS  245 

temperature  of  the  feed-water  will  always  be  less  than  the  temper- 
ature of  the  exhaust  steam.  (2)  Ordinarily,  the  pressure  in 
an  open  heater  is  but  slightly  in  excess  of  atmospheric  pressure. 
Ordinarily,  the  pressure  of  the  water  in  a  closed  heater  is  some- 
what in  excess  of  boiler  pressure.  (3)  An  open  heater  is  liable 
to  rupture  by  the  building  up  of  a  back-pressure,  due  to  sticking 
of  the  back-pressure  valves.  A  closed  heater  is  built  to  with- 
stand any  pressure  which  is  likely  to  occur.  (4)  Oil  in  the  exhaust 
steam  may  contaminate  the  feed-water  in  an  open  heater.  Oil 
cannot  enter  the  feed-water  from  the  exhaust  steam  in  a  closed 
heater.  (5)  Scale  and  other  impurities  precipitated  in  an  open 
heater  are  readily  removed.  It  is  difficult  to  remove  scale  from  a 
closed  heater.  If  the  feed  water  contains  a  high  content  of 
scale-forming  impurities,  then,  usually,  the  open  heater  is  the 
preferable  and  in  some  cases  the  only  permissible  type.  (6) 
An  open  heater  must  be  located  above  the  pump  suction.  The 
feed  pump  must  be  between  the  heater  and  the  boiler.  A  closed 
heater  may  be  located  anywhere  between  the  feed  pump  and  the 
boilers.  (7)  Where  the  water  is  taken  from  a  natural  source  of 
supply,  two  pumps  are  necessary  with  an  open  heater.  With  a 
closed  heater  only  one  pump  is  required  in  any  case.  (8)  With 
an  open  heater  the  feed  pump  handles  hot  water.  With  a  closed 
heater  the  'feed  pump  handles  cool  water.  (9)  An  open  heater 
cannot  be  installed  in  the  exhaust  line  from  a  condensing  engine 
as  can  a  closed  heater.  (10)  The  returns  from  a  heating  system 
cannot  be  delivered  directly  to  a  closed  heater  as  to  an  open 
heater. 

281.  In  The  Installation  Of  An  Open  Feed-Water  Heater 
the  following  general  directions  should  be  observed. 

DIRECTIONS. — (1)  Locate  the  heater  so  that  it  may  be  conveniently 
piped  to  the  source  at  the  exhaust-steam  supply  and  will  be,  at  the  same 
time,  as  close  as  possible  to  the  boilers. 

(2)  Set  the  heater  plumb  on  a  substantial  foundation  (Fig.  230)  of 
proper  height  to  bring  the  hot-water  outlet  to  the  feed-pump  at  least 
4  ft.  above  the  discharge-valve  deck  of  the  pump. 

(3)  Locate  the  feed-pump  (Fig.  228)  as  close  as  possible  to  the  heater. 
Also,  run  the  suction  pipe,  of  a  size  equal  to  the  outlet  orifice  of  the 
heater,  as  directly  as  possible.     If  the  pump  must  be  located  at  some 
distance  from  the  heater,  or  the  suction  connection  must  be  made  with 
a  number  of  sharp  turns,  either  the  suction  pipe  should  be  of  larger 


246  STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 

size  than  the  outlet  orifice  of  the  heater  or  the  heater  should  be  set  at  a 
greater  height  than  4  ft.  above  the  discharge-valve  deck  of  the  pump. 
In  some  cases  both  of  these  alternatives  may  be  desirable. 

(4)  Before  connecting  the  mechanism  of  the  float  (H,  Fig.  235)  to  the 
controlling  valve  in  the  water-supply  pipe,  see  that  the  float  and  mechan- 
ism move  freely. 

(5)  Pack  the  filtering  material,  excelsior  and  coke  (Fig.  235)  closely 
between  the  filter  plates. 

(6)  If  the  heater  is  to  be  connected  up  for  thoroughfare  service  (Fig. 
224)  attach  the  engine  exhaust  pipe  directly  to  the  heater  exhaust  inlet, 
and,  from  the  heater  exhaust  outlet,  run  a  pipe  to  the  atmosphere.     This 
pipe  should  be  of  the  same  size  as  the  exhaust  outlet.     A  back-pressure 
or  exhaust-relief  valve  should  be  placed  in  it  at  a  point  somewhere  beyond 
any  branch  connection  which  may  be  made  for  supplying  a  heating 
system,  or  for  other  purposes. 

(7)  See  that  the  back-pressure  valve  (V,  Fig.  229)  is  set  for  a  pressure 
not  higher  than  that  which  the  heater  will  safely  sustain. 

(8)  If  the  heater  is  to  be  connected  up  for  induction  service  (Fig.  228) 
run  a  branch  from  the  main  exhaust  pipe  to  the  heater  exhaust  inlet. 
This  branch  should  be  of  the  same  size  as  the  inlet  orifice  of  the  heater. 
It  should  contain  a  gate  valve,  so  that  the  heater  may  be  cut  out  for 
cleaning,  and  also  to  provide  a  means  for  regulating  the  supply  of  exhaust 
steam  delivered  to  the  heater. 

(9)  A  vent  pipe  (V,  Fig.  230)  should  be  attached  to  the  top  of  an  induc- 
tion heater.     This  is  to  allow  air  to  escape  and  to  insure  admission  of  the 
requisite  quantity  of  steam.     The  vent  pipe  may  be  screwed  into  a  reducer 
flange  bolted  to  the  heater  exhaust  outlet.     It  should  have  a  free  opening 
throughout  its  length.     The  valve,  V,  (Fig.  230)  in  the  vent  pipe  should 
never  be  closed  except  when  the  heater  is  cut  out  of  service  for  cleaning. 

(10)  Place  a  gate  valve  in  the  cold  water  supply  pipe  (F,  Fig.  235)  just 
beyond  the  controlling  valve.     Also  place  a  gate  valve  in  the  pump  suc- 
tion pipe. 

(11)  Connect  the  oil-drip  (W,  Fig.  235)  and  the  blow-off  pipe  to  the 
sewer  independently  of  each  other.     (If  it  is  desired  to  recover  and  filter 
the  oil  for  further  use,  the  oil-drip  may  be  piped  to  a  separate  reservoir.) 

(12)  Cover  the  heater  with  asbestos,  magnesia,  or  other  heat-insulating 
substance,  to  prevent  radiation  of  heat  therefrom. 

282.  In  The  Operation  Of  An  Open  Feed-Water  Heater 

the  following  general  directions  should  be  observed. 

DIRECTIONS. — (1)  When  the  heater  is  first  put  in  service,  the  cold-water 
controlling  valve  should  be  blocked  open,  also  the  blow-off  valve  should 
be  opened,  and  a  current  of  water  permitted  to  run  through  until  the 
heater  is  thoroughly  flushed  out.  The  blocking  may  then  be  removed 
from  the  controlling  valve  and  the  blow-off  valve  closed. 

(2)  The  lengths  of  the  float  connections  should  be  so  adjusted  that  the 


SEC.  283]  FEED-WATER  HEATERS  247 

controlling  valve  will,  respectively,  be  open  fully  and  closed  tightly  at  the 
predetermined  low  and  high  water  levels. 

(3)  The  blow-off  valve  should  be  opened  once  a  day  to  blow  out  the 
sediment  which  may  have  collected  in  the  bottom  of  the  heater. 

(4)  About  once  a  week,  or  oftener,  if  necessary,  the  coke  filter  bed 
should  be  flushed  out  by  opening  the  blow-off  valve  and  admitting  water 
under  pressure  through  the  flushing  pipe  (B,  Fig.  235). 

(5)  The  pans  of  an  open  heater  of  the  type  shown  in  Fig.  239  should  be 
removed  and  cleaned  whenever  the  depth  of  scale  is  sufficient  to  interfere 
with  their  operation  as  settling  basins.     The  time  allowable  before  this  is 
necessary  depends  on  the  nature  of  the  water. 

283.  In  The  Installation  and  Operation  Of  A  Closed  Feed- 
Water  Heater   the   following   general   directions   should   be 
observed : 

DIRECTIONS. — (1)  The  heater  should  be  connected  to  the  main  ex- 
haust pipe  as  near  the  engine  as  may  be  practicable. 

(2)  All  feed-water  and   blow-off  connections  should   be   made  with 
either  box  or  flange  unions,  so  that  the  parts  can  be  easily  taken  apart  for 
inspection. 

(3)  A  straightway  valve  or  plug  cock  should  be  inserted  in  the  blow-off 
pipe. 

(4)  The  safety  valve  on  the  feed-pipe,  in  the  case  of  a  water-tube 
heater,  or  on  the  heater  itself,  in  the  case  of  a  steam-tube  heater,  should 
be  loaded  from  15  to  20  Ib.  per  sq.  in.  above  the  boiler  pressure.     No 
other  valve,  of  any  kind,  should  be  placed  between  the  safety  valve  and 
heater. 

(5)  The  drip  pipes  should  be  of  the  same  size  as  the  drain  orifices  in 
the  heater.     The  drip  pipes  should  contain  as  few  bends  as  possible  and 
should  incline  downwards  from  the  heater  in  all  parts  of  their  lengths. 

(6)  The  heater  should  be  covered  with  a  heat  insulating  material  to 
prevent  loss  of  heat  by  radiation. 

(7)  The  blow-off  valve  should  be  opened  once  a  day  to  relieve  the 
heater  of  any  sediment  that  may  have  collected. 

(8)  When  the  plant  is  shut  down  in  cold  weather,  the  heater  should  be 
thoroughly  drained,  to  obviate  danger  of  freezing. 

284.  If  The  Safety  Valve  of  a  Closed  Feed-Water  Heater 
Will  Not  Remain  Tight  under  the  normal  operating  pressure 
it  should  be  examined  carefully  to  determine  the  cause.     If 
the  valve  is  of  the  lever  type,  extra  weights  should  not  be 
added  to  it  in  an  effort  to  make  it  tight.     Disaster  may  result 
from   such   procedure.     Neither   should   the   tension   of  the 
spring  be  increased,  if  it  is  of  the  spring  type,  unless  an  in- 
vestigation shows  that  the  spring-tension  is  too  low.     If  the 


248 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  7 


valve  does  not  close  tightly  after  blowing  off  or  if  it  "  simmers" 
instead  of  blowing,  it  usually  means  that  the  seat  or  valve 
is  in  bad  condition  or  that  the  adjusting  ring  is  so  far  from 
the  proper  position  that  the  valve  is  "out  of  control." 

285.  To  Get  The  Most  Effective  Service  From  A  Feed- 
Water  Heater,  It  Must  Be  Cleaned  at  regular  and  frequent 
intervals.     Local   conditions  must,  in  every  case,  determine 
the  frequency  of  the  cleanings.     But  in  no  case  should  the 
heater  be  operated  longer  than  a  month  without  cleaning. 

286.  Live-Steam   Heaters   And   Purifiers    (Fig.    239)    are 
intended,    primarily,    to    purify    the    feed-water.     They    are 


Lire  Steam  Pur/ f/'er-^ 


t  Steam-Supply  To  P\trhp 
•Steam  Supply  To  Pump  Through  Pur/f/er 
•Gravity  Feec/Fmm  Purifkr  To  &oi/ers 


-Open  Exhaust-    -^ucf 
Steam  Heater     fo  Pump 


•feed  Pump    "  ^By-Pass  For  Direct  Feed  fo  Boilers 


FIG.  252. — Hoppes  Live-Steam  Purifier  Installed  In  Connection  With  Exhaust-Steam 
Heater.  (When  the  purifier  is  in  operation,  the  pump  is  supplied  with  steam 
through  connection  F  in  order  that  air  and  non-condensable  gases  liberated  from  the 
feed- water  may  be  removed  from  the  purifier.) 

installed  (Fig.  252)  where  the  feed-water  contains  scale- 
forming  elements,  as  the  sulphates  of  lime  and  magnesia, 
which  precipitate  at  much  higher  temperatures  than  are 
obtainable  in  open  exhaust-steam  heaters. 

NOTE. — ALL  OF  THE  SCALE-FORMING  IMPURITIES  DISSOLVED  IN  A 
FEED- WATER  MAY  USUALLY  BE  PRECIPITATED  IN  A  LIVE-STEAM  PURI- 
FIER if  the  water  is  properly  pre-heated;  see  the  author's  STEAM  BOILERS. 
The  sulphates  of  lime  and  magnesia  precipitate  at  temperatures  above 
250  deg.  fahr.  The  carbonates  precipitate  at  a  much  lower  temperature. 
It  is  claimed  that  80  per  cent,  of  the  sulphates  will,  at  a  temperature  of 


SEC.  286]  FEED-WATER  HEATERS  249 

250  deg.  fahr.,  be  deposited  in  a  live-steam  purifier.  Also,  that  at  a 
temperature  of  300  deg.  fahr.,  all  of  the  sulphates  will  be  deposited  in  the 
purifier. 

QUESTIONS  ON  DIVISION  7 

1.  What  are  the  three  principal  reasons  why  a  feed-water  heater  should  be  used? 

2.  Give  an  approximate  rule  for  estimating  the  saving  due  to  heating  feed-water. 
Give  an  approximate  rule  for  estimating  the  increase  in  boiler  capacity  due  to  feed- 
water  heating. 

3.  How  does  a  feed-water  heater  protect  a  boiler  from  undue  strains  in  the  seams? 
Give  an  estimated  value  for  heat  strains  in  boiler  plates  caused  by  cold  water  in  a 
boiler. 

4.  What  is   an   open  heater?     A   closed  heater?     An   atmospheric  heater?     A   vacuum 
heater? 

5.  What  kind  of  heaters  are  used  as  primary  heaters?     How  are  they  connected  to 
other  equipment?     What  are  the  approximate  temperatures  in  a  primary  heater  with 
good  condenser  action?     What  condition  of  the  condenser  and  cooling  water  makes  the 
use  of  a  primary  heater  advisable? 

6.  What  is  a  secondary  heater?     How  is  it  connected  to  the  primary  heater.     What 
are  the  average  temperatures  for  an  open  atmospheric  heater  steam  supply  and  water 
outlet? 

7.  What  is  an  induction  heater?     A  through  heater?     How  is  each  piped? 

8.  What  operating  condition  governs  the  temperature  of  the  exhaust  steam  available 
for  use  in  a  heater? 

9.  Why  is  the  feed-water  temperature  obtained  with  a  closed  heater  ordinarily  lower 
than  that  obtained  with  an  open  heater? 

10.  What  functions  are  performed  by  an  ordinary  exhaust-steam  feed-water  heater? 
Describe  the  operation  of  an  open  heater  in  detail. 

11.  Name  several  scale-forming  impurities   that   may   be  precipitated  in  an   open 
heater. 

12.  Why  is  all  oil  objectionable  in  feed-water?     Why  is  cheap  oil  likely  to  be  especially 
objectionable? 

13.  What  common  dissolved  gases  are  objectionable  in  feed  water?     Why? 

14.  What   is   a   water-tube   heater?     A   steam-tube   heater?     A    parallel-current   heater? 
A  counter-current  heater?     A  single-flow  heater?     A  multiflow  heater? 

_  15.  What  is  a  coil  heater?     A.  film  heater? 

16.  What  advantages  are  claimed  for  spirally  corrugated  heater-tubes? 

17.  For  what  classes  of  service  are  steam-tube  heaters  particularly  adapted?     Why? 

18.  What  is  the  basis  of  the  heater  horsepower? 

19.  What  is  a  double-heater  installation? 

20.  What  are  the  relative  advantages  and  disadvantages  of  open  and  closed  heaters? 

21.  What  is  a  live-steam  purifier? 

22.  What  per  cent,  of  the  exhaust  steam  from  a  non-condensing  engine  does  a  feed- 
water  heater  ordinarily  consume  in  heating  the  feed  for  the  engine  boilers. 

23.  How  does  the  volume  of  an  open  feed- water  heater  affect  its  efficiency  as  a  purifier? 

24.  Give  a  few  general  directions  for  the  installation  of  an  open  feed-water  heater. 

25.  Give  a  few  directions  for  preparing  an  open  heater  for  service  and  keeping  it 
working  properly. 

26.  Why  should  an  oil  separator  usually  be  installed  in  the  steam  line  to  an  open 
heater? 

PROBLEMS  ON  DIVISION  7 

1.  Water  at  a  temperature  of  90  deg.  fahr.  is  available  for  feeding  the  boilers  in  a 
power  plant.  The  main  engine  runs  condensing.  It  develops  500  h.p.  on  a  steam- 
consumption  of  20  Ib.  per  h.p.  per  hr.  The  steam  consumption  of  the  auxiliaries  is 
about  11  per  cent,  of  that  of  the  main  engine.  If  the  exhaust  from  the  auxiliaries  is 
condensed  in  an  open  atmospheric  heater,  what  will  be  the  temperature  of  the  feed- 
water  as  delivered  to  the  boilers? 


250  STEAM  POWER  PLANT  AUXILIARIES  [Div.  7 

2.  A  boiler  generates  steam  at  a  pressure  of  150  Ib.  per  sq.  in.,  gage.     The  water  which 
is  fed  to  the  boiler  is  preheated  with  exhaust  steam  from  60  deg.  fahr.  to  210  deg.  fahr. 
What  saving  of  fuel  results  from  thus  utilizing  the  exhaust  steam? 

3.  The  coal  consumption  of  a  set  of  boilers  is  5  tons  per  day.     The  feed-water  is 
delivered  at  a  temperature  of  150  deg.  fahr.     It  is  estimated  that  by  using  a  quantity 
of  exhaust  steam  which  is  now  going  to  waste,  the  feed-water  may  be  delivered  at  a  tem- 
perature of  212  deg.  fahr.     The  average  steam  pressure  is  125  Ib.  per  sq.  in.,  gage.     The 
fuel  costs  $3.50  per  ton.     The  plant  operates  300  days  per  year.     It  will  cost  about 
$300  to  improve  the  present  heating  equipment.     The  rate  of  interest  on  the  invest- 
ment is  6  per  cent,  per  annum.     The  assumed  rate  of  depreciation  is  6.0  per  cent,  per 
annum.     It  will  probably  cost  $4  per  month  to  maintain  and  operate  the  apparatus. 
What  will  be  the  probable  annual  net  saving? 

4.  A  closed  exhaust-steam  feed-water  heater  is  required  to  heat  15,000  Ib.  of  feed 
water  per  hr.  from  70  to  200  deg.  fahr.  with  steam  at  220  deg.  fahr.     The  heater  is  to 
be  of  the  multiflow  plain  brass  water-tube  type.     What  should  be  the  area  of  the  tubing? 

5.  If  an  open  heater  heats  15,000  Ib.  per  hr.  of  feed-water  from  40  deg.  fahr.  to  205 
deg.  fahr.  with  steam  at  212  deg.  fahr.,  what  weight  of  steam  does  it  condense? 


DIVISION  8 
FUEL  ECONOMIZERS 

287.  A  Fuel  Economizer  (Fig.  253)  is  an  apparatus  in  which 
boiler  feed-water  is  preheated  by  the  combustion  gases  (Table 
288)  which  are  discharged  from  boiler-settings.  The  econo- 
mizer is  interposed  in  the  path  of  the  gases  between  the  boiler 
and  the  chimney. 


#»Vers--v2 


Fia.  253. — An  Economizer  Functions  To  Raise  The  Temperature  Of  The  Feed  Water. 
(Sturtevant  Economizer  Co.) 

288.  Table  Showing  The  Percentage  Of  The  Heat  Of 
The  Fuel  Which  Is  Present  In  The  Gases  Of  Combustion  As 
They  Leave  The  Boiler  (Green  Economizer  Co.). — Column  A 
is  based  on  an  air  supply  of  18  lb.,  per  pound  of  combustible. 
This  represents  average  underfeed  stoker  operation  with 
forced  draft.  Column  B  is  based  on  an  air  supply  of  24  lb., 
per  pound  of  combustible.  This  represents  average  overfeed 
or  natural-draft  stoker  operation.  Column  C  is  based  on  an 
air  supply  of  30  lb.,  per  pound  of  combustible.  This  represents 
average  operation  with  hand  firing  and  natural  draft. 

251 


252 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


Flue-gas  temperature 
in  deg.  fahr. 

Per  cent,  of  heat  of  fuel  in  flue  gases 

A 

B 

C 

300 

12.4 

350 

.... 

12.0 

14.9 

400 

.... 

14.0 

17.4 

450 

12.2 

16.1 

20.0 

500 

13.8 

18.2 

22.6 

550 

15.4 

20.3 

25.2 

600 

17.0 

22.4 

27.8 

650 

18.5 

24.4 

30.4 

700 

20.1 

26.5 

750 

21.7 

800 

23.2 

NOTE. — THE  HEAT  WHICH  Is  UTILIZED  IN  AN  ECONOMIZER  DOES 
NOT  REPRESENT  A  CLEAR  GAIN  (Fig.  254).  To  compensate  for  the  loss 
of  natural  draft,  which  results  from  lowering  the  chimney  temperature, 
it  is  generally  necessary  to  install  a  system  of  artificial  draft.  This 


...  .. 

Thls  Loss  Maybe 
Diminished  by  •:•  t 
Economizer  Service] 


FIG.   254. — Chart  Showing  Losses  In  Power  Plant  Operation. 

entails  an  extra  expense  for  draft  equipment,  and  for  the  subsequent 
operation  and  maintenance  thereof.  However,  it  is  often  profitable 
to  install  an  economizer  in  a  plant  of  greater  capacity  than  about  500 
boiler  horse  power. 

289.  There  Are  Two  General  Types  Of  Fuel  Economizers : 

(1)  The  independent  type  (Figs.  255  and  256).  (2)  The  integral 
type  (Figs.  257  and  258).  The  first  is  located  apart  from  the 


SEC.  289] 


FUEL  ECONOMIZERS 


253 


Feed-Wafer   Reversing         f«»y-.v.., W>ms         /Driving  Shaft  for  Scra 

Wq&tfMKticmsm  Jg&...)gfr.  flfr       \  Ob    jgft    ^^, 


FIG.  255. — An  Independent  Fuel-Economizer.     (Green  Economizer  Co.) 


U,...c,.,,        p,,..  .,^       ^..~..~,       r---,r,^ 
, — [>?':>  :  -V^~l  .  -  -*•;;  ^V-'-'?'.  |j.  ±± '.-'---.--'.<  -||.'r'-;--'.-.-'-:?H  . 


.•Induced  Draft 
Fan 


FIG.  256. — A    Typical    Installation    Of    Independent    Fuel-Economizers.     (Hamptoi 
Mills,  East  Hampton,  Mass.) 


254  STEAM  POWER  PLANT  AUXILIARIES  [Div.  8 

boiler  setting.  The  second  is  located  within  the  boiler  setting. 
Thus,  it  practically  forms  an  integral  part  of  the  boiler 
structure. 


FIG.  257. — High-  And  Low-Pressure  Economizers.     (Kansas  City  Light  And  Power  Co. ) 

290.  An  Independent  Economizer  (Fig.  255)  consists, 
essentially,  of  a  double  series  of  cast-iron  headers,  or  mani- 
folds (Fig.  259)  which  are  connected  together  by  vertical 
tubes.  The  tubes  are  commonly  made  of  cast-iron.  Their 
usual  dimensions  are  4%6-in.  diameter  and  9-  to  12-ft.  length. 


SEC.  291] 


FUEL  ECONOMIZERS 


255 


The  water,  which  is  discharged  by  the  boiler  feed-pump, 
passes  through  the  headers  and  tubes  of  the  economizer  before 
it  enters  the  boiler.  The  hot  gases,  which  flow  from  the 
boiler-setting  to  the  chimney,  pass  (Figs.  260  and  261)  through 
the  spaces  between  the  economizer  tubes.  The  heat  in  the 
gases  is  thereby  transmitted  to  the  feed-water. 


Combustion    Chamber 
Arranged         For 
Oil       Firing 


FIG.  258.— Badenhausen   Boiler   Directly   Connected   With   Integral   Economizer    Or 

Preheater. 

NOTE. — ECONOMIZER-TUBES  MAY  BE  ARRANGED  IN  EITHER  STRAIGHT 
OR  STAGGERED  Rows.  The  staggered  arrangement  (Fig.  261)  affords 
the  greater  facility  for  heat- transfer  from  the  gases.  The  straight 
arrangement  (Fig.  260)  offers  the  least  obstruction  to  the  draft.  Thus 
the  advantage  of  either  arrangement  is  apparently  offset  by  the  dis- 
advantage of  the  other. 

291.  Integral  Economizers  are  designed  to  withstand  either 
high  pressures  or  low  pressures.  High-pressure  integral  econo- 
mizers are  so  located  (Figs.  257  and  258)  as  to  receive  the 
gases  directly  as  they  issue  from  contact  with  the  boiler  sur- 


256 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  S 


Drotw-Out 


/////////^ 


Fia.  259. — Construction    Of    Headers    And    Tubes    Of   Sturtevant    Economizer    And 
Method  Of  Removal  And  Replacing  Of  Tubes. 


Direction  of  das  flow-- 

FIQ.    260. — Economizer    Tubes    In 
Straight  Rows. 


Tubes-. 


Direction  of  6«5  FtoHf* 

Fio.    261. — Economizer    Tubes    In 
Staggered  Rows. 


SEC.  292]  FUEL  ECONOMIZERS  257 

faces.  These  economizers  are,  therefore,  built  with  wrought 
iron  or  steel  tubes  and  drums.  Low-pressure  integral  econo- 
mizers are  so  located  (Fig.  257)  as  to  receive  the  gases  at  a 
comparatively  low  temperature.  Hence,  these  economizers 
are  usually  built,  similarly  to  the  independent  type  of  econo- 
mizer (Sec.  289),  with  cast-iron  tubes  and  headers. 

292.  Certain  Advantages  And  Disadvantages  Attend  The 
Use  Of  Cast-iron,  Wrought  Iron  And  Steel  In  Economizer 
Construction  (Sees.  290  and  291). — Cast-iron  tubes  and  headers 
are  less  susceptible  to  corrosion  than  are  those  which  are  made 
of  wrought  iron  or  steel.     But  the  liability  of  cast-iron  tubes 
and  headers  to  fail  under  the  stresses  of  expansion  and  con- 
traction, and  pressure  is  by  far  the  greater. 

NOTE. — CORROSION  OF  ECONOMIZERS  may  be  due,  internally,  to  an 
acid  property  of  the  feed-water.  Externally  it  may  be  due  to  sulphurous 
acid  or  dilute  sulphuric  acid  which  are  formed  by  the  action  of  moisture 
and  SC>2  in  the  sooty  deposits  on  the  tubes.  The  moisture  may  come 
from  leaky  joints,  or  it  may  be  due  to  a  sweated  condition  of  the  tubes. 

SWEATING  OF  ECONOMIZER-TUBES  occurs  when  the  temperature  of  the 
metal  falls  below  the  dew-point  of  the  combustion  gases.  This  condition 
will  generally  result  when  water  at  a  temperature  less  than  about  130  deg. 
fahr.  is  pumped  through  the  economizer.  Certain  economizer  manufac- 
turers recommend  that  the  entering  feed-water  temperature  should 
be  at  least  90  to  100  deg.  fahr.  If  the  available  feed  water  is  colder,  a 
by-pass  may  be  arranged  to  pass  some  hot  water  into  the  feed  line. 

293.  Cleanliness  Of  The  Tube-Surfaces,  Both  Inside  and 
Outside,  Is  Essential  To  The  Effectiveness  Of  A  Fuel  Econo- 
mizer.— The    soot    which  is  mingled  with  combustion-gases 
adheres  very  readily  to  economizer-tubes.     This  is  due  to  the 
comparatively  low  temperature  of  the  tubes.     Soot  is  an  excep- 
tionally poor  heat-conductor.     Hence  the  urgent  necessity  for 
its  removal  from  the  tubes  is  apparent. 

294.  Two  Methods  Are  Available  For  Removing  Soot  From 
Economizer -Tubes :  (1)  Scraping.     (2)  Blowing.     The  scrap- 
ing-method (Figs.  255  and  262)  is  the  more  commonly  used. 

EXPLANATION. — Economizer-tube  scrapers  (Fig.  263)  are  in  the  form  of 
sleeves  which  encircle  the  tubes.  These  sleeves  are  caused  to  traverse 
the  tubes,  from  end  to  end,  by  means  (Figs.  255  and  262)  of  a  geared 
mechanism.  The  soot,  which  is  scraped  off  by  the  beveled  edges  of  the 
sleeves,  falls  into  a  pit  beneath  the  economizer.  It  is  then  removed 
17 


258 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


through  clean-out  doors.     Or,  the  soot  may  drop  into  a  pit  (Fig.  255) 
whence  it  is  conveyed  away  through  a  pipe. 

NOTES. — ECONOMIZER  SOOT-BLOWERS  (Fig.  264)  are  of  the  same  type 
as  those  which  are  used  with  water-tube  boilers.     These  blowers  are 


.-•Driving 
{   Pulley 


FIG.  262. — Mechanism  For  Automatic  Reversal 
Of  Travel  Of  Soot-Scrapers  On  "Green"  Fuel- 
Economizers.  Bevel-Gear  Pinions  Bi  And  82 
Are  Loose  On  Shaft  S. 


FIG.    263.— Soot-Scraper    For 
Economizer-Tubes. 


described  in  the  author's  STEAM  BOILERS.  It  is  claimed  that  they 
remove  the  soot  entirely  from  the  tube-surfaces.  With  the  use  of 
sleeve-scrapers  (Fig.  263)  a  thin,  compact,  film  of  soot  may  constantly 
remain  on  the  surfaces. 


,-rtow  of  Gases 


gprpcfreV  Blower  Elements-.. 


Automatic  Drain--' 
I  Drain  Valve--' 
^~ 


FIG.  264.— "Green"  Fuel-Economizer  Equipped  With  Vulcan  Soot  Blowers. 

THE  POWER  EXPENDED  IN  THE  OPERATION  OF  ECONOMIZER-TUBE 
SCRAPERS  may  be  approximately  1  h.p.  per  1000  sq.  ft.  of  economizer 
surface. 

THE  STEAM-CONSUMPTION  OF  A  SOOT-BLOWING  SYSTEM  depends  upon 
the  size  of  the  system  and  the  time-interval  during  which  it  must  be  used 


SEC.  295] 


FUEL  ECONOMIZERS 


259 


to  effectually  remove  the  soot.  A  system  consisting  of  six  blower-units, 
each  fitted  with  38  nozzles,  will  consume  2600  Ib.  of  steam  during  a  blow- 
ing period  of  six  minutes  (POWER  HOUSE;  July  5,  1919,  p.  272). 

295.  Deposits  Of  Scale  And  Sediment  In  Economizer -Tubes 
Are  Detrimental  To  Economy  (Fig.  265).— Where  the  feed- 
water  contains  scale-  and  mud-forming  impurities,  the  econo- 
mizer should  be  frequently  blown  down.  Also,  the  tubes 
should  be  washed  out,  as  often  as  is  necessary,  with  a  hose. 
Formation  of  hard  scale  may,  by  these  means,  be  prevented. 


ON 

2  JO 

s 

^ 

, 

A 

??fi 

A 

\ 

v 

/ 

\ 

y 

*\ 

\ 

-TV, 

A' 

?  70 

s. 

*!* 

•-* 

""^ 

\ 

/ 

\, 

v* 

•—  . 

J 

r 

^i 

^ 

f 

>- 

*•-», 

% 

r 

->, 

—  -1 

^ 

s/ 

^ 
Rk 

/Ml 
Yin 

E 

51 

A 

J 

V 

x 

^— 

\ 

3 

/ 

- 

j 

V- 

B: 

700 

* 

2  3  4  5  6  7  8  9  10  II  12  13  14  15  16  IT  18  19  20  21  2223  24  25  26  27  28  29  30  31 
Days  of          Month 

FIG.  265. — Diagram  Showing  Daily  Average  Consumption  Of  Coal  When  Econ- 
omizer Tubes  Were  Clean  And  When  They  Were  Lined  With  Scale.  Graph  A-A 
Shows  Consumption  When  Tubes  Are  Lined  With  Scale.  A'-A'  Is  The  Average  Of 
A-A.  Graph  B-B  Results  When  Tubes  Are  Clean.  B'-B'  Is  The  Average  of  B-B. 

NOTE. — SCALE  DOES  NOT  FORM  As  READILY  IN  ECONOMIZERS  As 
IN  BOILERS.  This  is  due  to  the  lower  temperature  of  the  water  in 
economizers.  The  temperature  is,  however,  usually  high  enough  to 
cause  precipitation  of  sedimental  impurities.  The  comparatively-low 
velocity  of  the  water-flow  in  an  economizer  facilitates  such  precipitation. 
Hence  the  sediment  readily  settles  into  the  bottom  headers,  whence  it 
may  be  blown  out  through  the  blow-off  valves. 

296.  An  Economizer  Should  Be  Fitted  With  Instruments 
For  Showing  The  Combustion-Gas  And  Feed-Water  Tem- 
peratures (Table  301). — Thermometers  should  be  inserted  in 
the  feed-water  connections  to  the  economizer,  both  at  inlet 
and  outlet.  Also,  a  pyrometer  should  be  inserted,  at  each 
end  of  the  economizer,  in  the  path  of  the  combustion-gases. 
These  instruments  afford  a  ready  means  for  checking  the 
performance  of  the  economizer. 

EXPLANATION. — Suppose  the  instruments  were  to  show  a  steady 
increase,  above  normal,  of  the  flue-gas  temperature  at  exit  from  the 


260 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


economizer,  while  the  temperature  of  the  outgoing  feed  water  steadily 
decreases.  This  condition  would  probably  indicate  that  the  heat  is 
excluded,  by  a  steadily  increasing  coating  of  soot,  from  the  tube  surfaces. 

297.  Infiltration  Of  Air  Through  The  Setting  Of  An  Econo- 
mizer Is  Detrimental  To  Economy. — Cool  air,  passing  in  through 
crevices  in  the  setting,  will  mingle  with  the  current  of  combus- 
tion gases.     The  air  will  thereby  absorb  heat  from  the  gases. 
Hence  the  quantity  of  heat  delivered  to  the  water,  flowing 
through  the  economizer,  will  be  diminished. 

NOTE. — Leakage  of  air  into  an  economizer  setting  may  occur  where  the 
tubes  are  cleaned  with  scrapers.  The  openings  through  which  the 
scraper-chains  pass  may  afford  ready  ingress  for  air.  This  difficulty  does 
not  attend  the  use  of  blowers. 

298.  Excessive  Leakage  Of  Air  Into  An  Economizer  Setting 
May  Be  Detected  By  Observing  The  CO2  Drop  Through  The 
Economizer. — A  drop  of  about  2  per  cent,  may  reasonably  be 


3     1 
5   10 


w     10  20  30  40  50  60  10  80  90  100  110  120  130 
Lbof  Waste  Gas  per  Lb.  of  Coat 

FIG.  266. — Chart  Showing  The  founds  Of  Gas  Per  Pound  Of  Illinois  Coal  Corresponding 
To  Percentages  Of  COa.    (Power  Plant  Engineering,  Apr.  1,  1919.) 

expected.     When  this  percentage  of  drop  is  exceeded,  the 
leakage  of  air  is  probably  excessive. 

EXAMPLE. — The  combustion-gases,  passing  from  a  boiler,  have  a 
temperature  of  600  deg.  fahr.  and  contain  10  per  cent,  of  CO2  as  they 
enter  an  economizer.  As  they  leave  the  economizer,  due  to  infiltration 
of  air,  the  gases  have  a  temperature  of  300  deg.  fahr.  and  contain  6  per 
cent,  of  CO2.  The  outside-air  temperature  is  70  deg.  fahr.  The  specific 


SEC.  298] 


FUEL  ECONOMIZERS 


261 


262 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


heat  of  the  .gases  =  0.24.  It  is  assumed  (Fig.  266)  that  1  Ib.  of  coal 
yields  15.5  Ib.  of  combustion  gases  when  the  COg  amounts  to  10  per  cent., 
and  26  Ib.  of  gases  when  the  CO2  amounts  to  6  per  cent.  What  is  the 
percentage  of  heat-loss? 

SOLUTION. — The  heat,  above  70  deg.  fahr.,  which  is  contained  in  the  gases 
as  they  enter  the  economizer  =  (600  -  70)  X  15.5  X  0.24  =  1,971.6 
B.t.u.  per  Ib.  of  coal  burned  on  the  grate.  The  infiltration  of  air  amounts  to 
26  —  15.5  =  10.5  Ib.  per  Ib.  of  coal  burned.  The  heat  required  to  raise  the 
temperature  of  the  infiltered  air  to  300  deg.  fahr.  =  (300  -  70)  X  10.5  X 
0.24  =  579.6  B.t.u.  per  Ib.  of  coal  burned.  Hence,  the  percentage  of 
heat-loss  =  (579  -=-  1971.6)  X  100  =  29.4  per  cent. 

299.  The  Draft-Pressure  Drop  Through  An  Economizer 

(Fig.  267)  depends  upon  the  arrangement  (Sec.  290)  of  the 
economizer-tubes,  the  velocity  of  the  gases,  and,  perhaps, 


Economi: 


Draft,  Inches  of  Water 
0 15       0.20       a?5      CX30      035 


Uptake 


Economizer 
Entran 


1st  Pass 


2mLPois5 


Exit 


FIG.    268. — Draft    Pressure    Drop  FIG.  269. — Diagram  Showing  Arrange- 

Through    8,500  Sq.    Ft.,    3    Section  ment  Of  Forced  Draft  And  Induced  Draft 

Economizer-Fan  Draft.  Fans   In   Connection   With  Economizer. 

(B.  F.  Sturtevant  Co.) 

upon  conditions  peculiar  to  each  installation.  It  may  (Fig. 
268)  vary  from  0.15  to  more  than  0.3  in.  of  water  column. 
The  frictional  resistance  of  the  tubes  is  directly  proportional 
to  the  length  of  the  economizer  and  to  the  square  of  the 
velocity  of  the  gases. 

NOTE. — ECONOMIZERS  GENERALLY  PROVE  UNPROFITABLE  WHERE 
CHIMNEYS  ARE  ALONE  DEPENDED  UPON  To  CREATE  DRAFT  PRESSURE. 
Cooling  of  the  flue  gases  (Table  300)  by  the  economizer  diminishes  the 
draft-producing  effectiveness  of  the  gases.  In  addition  to  the  reduction 
of  natural  draft,  due  to  this  cause,  there  is  the  loss  of  draft-pressure  due 
to  pushing  the  gases  against  the  frictional  resistance  of  the  economizer. 


SEC.  300] 


FUEL  ECONOMIZERS 


263 


The  cost  of  the  additional  chimney-height,  necessary  to  compensate  for 
these  deficiencies,  will  often  more  than  offset  the  possible  gain  due  to 
heating  the  feed-water. 

ARTIFICIAL  DRAFT  Is  GENERALLY  USED  WITH  ECONOMIZER  INSTAL- 
LATIONS. The  draft  may  be  either  forced  (Fig.  269)  or  induced.  Sys- 
tems of  artificial  draft  are  illustrated  and  described  in  the  author's 
STEAM  BOILERS. 


300.  Table  Showing  Height  Of  Water  Column  Due  To 
Unbalanced  Pressures  In  Chimney  100  Feet  High.  Tempera- 
tures are  in  degrees  Fahrenheit. 


Temperature 
in 
chimney 

Temperature  of  external  air.     (Barometer  14.7  Ib.) 

0° 

10° 

20° 

30° 

40° 

50° 

60° 

70° 

80° 

90° 

100° 

200° 

.453 

.419 

.384 

.353 

.321 

.292 

.263 

.234 

.209 

.182 

.157 

220 

.488 

.453 

.419 

.388 

.355 

.326 

.298 

.269 

.244 

.217 

.192 

240 

.520 

.488 

.451 

.421 

.388 

.359 

.330 

.301 

.276 

.250 

.225 

260 

.555 

.528 

.484 

.453 

.420 

.392 

.363 

.334 

.309 

.282 

.257 

280 

.584 

.549 

.515 

.482 

.451 

.422 

.394 

.365 

.340 

.313 

.288 

300 

.611 

.576 

.541 

.511 

.478 

.449 

.420 

.392 

.367 

.340 

.315 

320 

.637 

.603 

.568 

.538 

.505 

.476 

.447 

.419 

.394 

.367 

.342 

340 

.662 

.638 

.593 

.563 

.530 

.501 

.472 

.443 

.419 

.392 

.367 

360 

.687 

.653 

.618 

.588 

.555 

.526 

.497 

.468 

.444 

.417 

.392 

380 

.710 

.676 

.641 

.611 

.578 

.549 

.520 

.492 

.467 

.440 

.415 

400 

.732 

.697 

.662 

.632 

.598 

.570 

.541 

.513 

.488 

.461 

.436 

420 

.753 

.718 

.684 

.653 

.620 

.591 

.563 

.534 

.509 

.482 

.457 

440 

.774 

.739 

.705 

.674 

.641 

.612 

.584 

.555 

.530 

.503 

.478 

460 

.793 

.758 

.724 

.694 

.660 

.632 

.603 

.574 

.549 

.522 

.497 

480 

.810 

.776 

.741 

.710 

.678 

.649 

.620 

.591 

.566 

.540 

.515 

500 

.829 

.791 

.760 

.730 

.697 

.669 

.639 

.610 

.586 

.559 

.534 

264 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


301.  Table  Of  Actual  Temperatures  Obtained  in  Typical 
Economizer  Installation  (Green  Economizer  Co.). 


Name  of  plant 


Temperatures 


Entering 
econo- 
mizer, 
deg. 
fahr. 


Leaving 
econo- 
mizer, 
deg. 
fahr. 


Of  water 


Entering 
econo- 
mizer, 
deg. 
fahr. 


Hollister  Mining  Company 598 

Mac  Sim  Bar  Paper  Co 630 

Mary  Charlotte  Mining  Co 500 

Kellogg  Toasted  Corn  Flake  Co 560 

Louisville  Water  Works 455 

Wessniger-Gaulbert  Realty  Co 510 

Galveston  Ice  Company 500 

Gilbert  Paper  Company • 500 

Bemis  Bros.  Bag  Co 550 

Great  Northern  Railway 600 

Portland  Railway  &  Light  Co 475 

Graniteville  Manufacturing  Co 442 

Arkwright  Mills 680 

Arnold  Print  Works 510 

Blackstone  Manufacturing  Co 436 

Champion  International  Co 800 

Granite  Mills  No.  1 455 

Hoosic  Cotton  Mills 430 

Kunhardt  Company 550 

Lancaster  Mills 655 

Nonotuc  Silk  Co 430 

Pierce  Manufacturing  Co 638 

Stanley  Works 700 

American  Thread  Co 455 

Lonsdale  Co 475 

American  Brass  Company 575 

Baltic  Mills 475 

Bridgeport  Malleable  Iron  Co 560 

Lawton  Mills 505 

Union  Metallic  Cartridge  Co 500 

Waterbury  Clock  Company 550 

American  Agricultural  Chemical  Co 415 

Chelsea  Mills  No.  1 460 

Remington  Salt  Co 700 

Saratoga  Victory  Manufacturing  Co 315 

Delaney  &  Company 670 

Bird  &  Son 600 

Berlin  Mills  Co 540 

Winchester  Repeating  Arms  Co 553 

Hammermill  Paper  Co 600 

Imperial  Steel  Company 386 


298 
418 
300 
320 
325 
325 
250 
320 
350 
420 
245 
263 
375 
280 
263 
570 
245 
275 
300 
270 
325 
434 
300 
345 
234 
425 
265 
300 
301 
350 
330 
265 
299 
500 
225 
450 
300 
217 
312 
300 
265 


202 
190 
210 
212 
164 
192 
100 
212 
180 
140 
130 
118 
120 
118 

77 
240 

82 
122 
100 
110 

36 
121 
150 
150 
101 
160 
120 
206 
140 
175 
130 

64 
160 
160 

75 
180 
130 

78 
194 
150 

95 


SEC.  302] 


FUEL  ECONOMIZERS 


265 


302.  The  Relative  Current-Flow  Of  The  Water  And  Gases 
Passing  Through  An  Economizer  may  be:  (1)  In  the  same 
direction.  (2)  In  opposite  directions.  The  first  is  called  a 
parallel-flow.  The  second  is  called  a  contra-flow.  For  a 
parallel-flow  (Fig.  270)  the  feed-water  and  the  combustion- 
gases  enter  the  economizer  at  the  same  end.  Thus  the  coolest 
part  of  the  water-current  abstracts  heat  from  the  hottest  part 
of  the  gas-current.  For  a  contra-flow  (Fig.  271)  the  feed- 
water  and  the  gases  enter  the  economizer  at  opposite  ends. 
Thus  the  water  is  first  heated  by  the  cooler  gases  and  later,  as 
it  passes  on  through  the  economizer,  by  the  hotter  gases.  A 
larger  transfer  of  heat  from  the  gases  to  the  water  occurs  with 


Feed-Water  Outlet^ 


Stack-' 


*- -Feed -Water  Inlet 


FIG.  270.  —  Illustration  Of  Water 
And  Gas  Flow  In  Parallel  Flow  Econ- 
omizer Installation. 


E  c  o  n  o    m 


Gases 
Wafer 


•Gas,  from 
Boiler 


Feed-Water.- --+ 
Inlet 


FIG.  271.— The    Flow   Of   Water   And" 
Gases  In  Counter  Flow  Economizer. 


a  contra-flow  than  with  a  parallel-flow.  This  is  due  to  the 
minimum  temperature-difference,  between  the  water  and  the 
gases,  being  (Fig.  272)  greater  with  a  contra-flow  than  (Fig. 
273)  with  a  parallel-flow. 


EXPLANATION. — Suppose  the  temperatures  of  the  combustion-gases 
and  feed-water,  at  entrance  to  an  economizer  which  is  arranged  for  a 
parallel-flow,  are,  respectively,  600  deg.  fahr.  and  100  deg.  fahr.  Also, 
suppose  the  temperature  of  the  gases  at  exit  from  the  economizer  to  be 
340  deg.  fahr.  Then  (Fig.  273),  if  the  water  is  to  be  heated  to  a  tempera- 
ture of  220  deg.  fahr.,  the  economizer  must  have  about  8,000  sq.  ft.  of 
heating-surface.  The  minimum  temperature  difference  =  340  -  220  = 
120  deg.  fahr. 

Now  suppose  the  economizer  to  be  arranged  for  a  contra-flow.  Then 
(Fig.  272),  7,300  sq.  ft.  of  heating-surface  would  suffice  to  produce  a  final 


266 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


feed-water  temperature  of  220  deg.  fahr.,  with  a  gas-temperature,  at 
exit,  of  340  deg.  fahr.  The  minimum  temperature  difference 
=340  - 100  =  204  deg.  fahr. 


—  490 


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Fio.  272. — Characteristics  Of  Econ- 
omizer Arranged  For  Contra-Flow  Of 
Gases  And  Water.  (Power,  Apr.  22, 
1919.) 


mmmw  mmmmmim 

So(.  Ft.  of  EconomizerSurface 

FIG.  273. — Characteristics  of  Econ- 
omizer When  Gases  And  Water  Flow 
In  The  Same  Direction.  (Power,  Apr. 
22,  1919.) 


303.  The  Ratio  Of  The  Loss  Of  Combustion-Gas  Tempera- 
ture To  Gain  Of  Feed  -Water  Temperature  In  An  Economizer 
may  be  found  by  the  following  formula: 


(82) 


(ratio) 


Wherein  X  =  ratio  of  decrease  of  gas-temperature  to  increase 
of  water-temperature.  T/a  =  loss  of  gas-temperature,  in 
degrees  Fahrenheit.  T/w  =  gain  of  water-temperature,  in 
degrees  Fahrenheit.  Cw  —  assumed  specific  heat  of  water  = 
1.0.  Cg  =  assumed  specific  heat  of  combustion-gases  =  0.24. 
Ww  =  weight  of  water  evaporated  in  boiler,  in  pounds  per 
pound  of  coal  burned.  Wg  =  weight  of  combustion-gases,  in 
pounds  per  pound  of  coal  burned. 

EXAMPLE.  —  An  average  of  13  Ib.  of  combustion-gases  are  produced, 
per  pound  of  coal  burned,  in  a  boiler-furnace.  An  average  of  6.5  Ib. 
of  water,  per  pound  of  coal  burned,  is  evaporated  in  the  boiler.  What 
is  the  ratio  between  the  gas-  and  water-temperature  changes  which  occur 
in  the  economizer?  SOLUTION.  —  By  For.  (82)  X  =  T/g/Tfv>  =CV>WV>/ 
CgWg  =  (1  X  6.5)  4-  (0.24  X  13;  =  2.08. 


SEC.  304] 


FUEL  ECONOMIZERS 


267 


NOTE. — The  ratio-value  obtained  in  the  solution  of  the  preceding 
example  is  commonly  assumed  to  represent,  approximately,  economizer- 
performance  in  general.  It  is  based  upon  the  assumption  that  infiltra- 
tion of  air  through  the  setting  (Sec.  297),  and  radiation  of  heat  from 
the  economizer,  are  both  reduced  to  a  minimum. 


304.  Additional  Heat- 
ing-Surface Obtained  By 
Installing  An  Economizer 
Will  Prove  More  Effect- 
ive, Than  Would  An 
Equivalent  Addition  Of 
Boiler  Heating-Surface, 
in  absorbing  heat  from 
the  combustion  -gases 
which  are  discharged  from 
a  set  of  boilers.  This  is 
due  to  the  greater  temper- 
ature-difference between 
the  water  and  the  gases. 


Fia.  274. — diagram  Showing  Gas  Tempera- 
tures In  A  Water-Tube  Boiler  Containing  6,000 
Sq.  Ft.  Of  Heating-Surface. 


Q   W 


EXPLANATION. — Suppose  the  steam-pressure  in  a  boiler  is  150  Ib.  per 
sq.  in.,  gage.  Then  the  temperature  of  the  water  in  the  boiler  will  be 
about  358  deg.  fahr.  Suppose  the  boiler  heating-surface  is  of  such 
extent  that  it  will  lower  the  combustion-gas  temperature  in  the  last 

pass  (Figs.  274  and  275)  to  500 
deg.  fahr.  Then  the  temperature- 
difference  between  the  inside  and 
the  outside  of  the  boiler  heating- 
surface  =  500  -  358  =  142  deg. 
fahr.  If  the  gases  were  now  to  flow 
in  contact  with  additional  boiler 
heating-surface,  the  rapidity  of  heat- 

K>    20    30    40    so    GO    TO    so    90   wo     transfer  from  the  gases  to  the  water 
Per  Cent  of  Boiler, Tube  Surface  Passed      woujd    (see  the  author's  PRACTICAL 

FIG.  275.-Graph  Showing  Gas  Tern-  HEAT)  be  in  dire.ct  Proportion  to 
peratures  In  A  Water-Tube  Boiler  Oper-  this  temperature-difference.  But  if 
ating  At  About  10  Sq.  Ft.  Of  Heating  the  gases  were  to  traverse  an  econ- 

Surface  Per  Boiler   Horse  Power.     (Gas     Omizer  through  which  water  at  150 

£&&&£}*   ^  1S   °°mPar"    deg.  fahr.  is  being  pumped,  then  the 

rapidity  of  heat-transfer  from   the 

gases   to  the  water  would  be  in   direct  proportion  to   a  temperature- 
difference  of  500  -  150  =  350  deg.  fahr. 


268 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  8 


305.  The  Least  Temperature -Difference  That  Can  Be 
Profitably  Permitted  Between  The  Inside  And  The  Outside 
Of  Boiler  Heating-Surface  may  be  determined,  approximately, 
from  the  chart  (Fig.  276)  which  has  been  computed  for  average 

conditions  of  boiler  service.  From 
the  temperature-difference  so  ob- 
tained, the  lowest  flue-gas  temper- 
ature may  be  computed.  The 
maximum  amount  of  heating- 
surface  which  a  boiler  should  have, 
for  given  conditions  of  operation, 
may  then  (Fig.  277)  be  determined. 


0    0 


Cost  of  Coal  in  Dollars  per  Ton 

FIG.  276. — Chart  Showing  The 
Least  Temperature  Difference  Be- 
tween The  Temperatures  Of  Gases 
And  Water  Under  Different  Con- 
ditions At  Which  Additional  Boiler 
Surface  Ceases  To  Pay  Dividends. 
(Green  Economizer  Co.) 


NOTE. — If  the  heating-surface  of  a 
boiler  is  too  extensive,  the  temperature- 
difference  in  the  last  pass  will  be  insuf- 
ficient to  insure  effective  heat-transfer. 

EXAMPLE. — A  set  of  boilers  is  to  deliver 
steam  at  a  pressure  of  200  Ib.  per  sq.  in.  The  daily  period  of  operation 
is  to  be  12  hr.  Coal  will  cost  $3  per  ton.  What  is  the  maximum 
amount  of  heating  surface,  consistent  with  economical  performance, 
which  each  boiler  should  have? 


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•     Boiler  Surface  in  Square  Feet  per  Boiler  Horse  Power  Developed 

FIG.  277. — Chart  Showing  Flue  Gas  Temperatures  Corresponding  To  Different 
Amounts  Of  Heating  Surface  Per  Boiler  Horse  Power  Developed.  Each  Point  Repre- 
sents A  Test.  (Green  Economizer  Co.) 

SOLUTION. — For  a  12-hr,  daily  run,  with  coal  at  $3  per  ton,  the  least 
temperature-difference,  consistent  with  profitable  operation  of  the  boiler, 
is  (Fig.  276)  200  deg.  fahr.  The  temperature  of  steam  at  a  pressure  of 
200  Ib.  per  sq.  in.,  gage,  is  about  388  deg.  fahr.  Hence,  the  lowest  tern- 


SEC.  306] 


FUEL  ECONOMIZERS 


269 


perature  of  the  combustion-gases  would  be  (200  +  388)  =  588  deg.  fahr. 
The  permissible  extent  of  heating-surface  is,  therefore,  (Fig.  277)  about  5.5 
sq.  ft.  per  boiler  h.p. 

306.  The  Least  Temperature -Difference  That  Can  Be 
Profitably  Permitted  Between  The  Inside  And  The  Outside 
Of  Economizer  Heating-Surface  may  be  determined,  ap- 
proximately, from  the  chart  (Fig.  278)  which  has  been  com- 
puted for  different  conditions  of 
economizer  service.  Local  con- 
ditions, peculiar  to  individual 
plants  may,  however,  sometimes 
affect  the  accuracy  of  the  deter- 
minations. 


Cost  of  Coal  in  Do II a  r  s  Per  Ton 


Fio.  278.— Chart  Showing  The 
Least  Temperature  Differences  Be- 
tween The  Temperature  Of  Gases 
And  Water,  Under  Different  Con- 
ditions, At  Which  Additional  Econ- 
omizer Surface  Ceases  To  Pay 
Dividends.  (Green  Economizer 
Co.) 


EXAMPLE. — A  contra-flow  economizer 
(Fig.  271)  is  to  be  installed  in  connection 
with  the  set  of  boilers  mentioned  (Sec. 
305)  in  the  preceding  example.  The  tem- 
perature of  the  feed-water,  at  entrance  to 
the  economizer,  is  to  be  150  deg.  fahr. 
The  temperature  of  the  entering  gases  is 
588  deg.  fahr.  What  should  be  the  least 
temperature,  consistent  with  economy, 
of  the  gases  issuing  from  the  economizer? 

SOLUTION. — For  a  12-hr,  daily  run,  with  coal  at  $3  per  ton,  the  least 
temperature-difference,  consistent  with  profitable  operation  of  the  econo- 
mizer, is  (Fig.  278)  90  deg.  fahr.  Hence,  the  exit-temperature  of  the 
gases  should  le  (90  +  150)  =  240  deg.  fahr. 

307.  The  Ratio  Of  Economizer  Heating-Surface  In  Square 
Feet  To  Boiler -Horsepower  usually  ranges  from  about  4:1 
to  8:1.     An  extent  of  heating-surface,  for  the  economizer  in 
the  above  example,  which  would  give  a  mean  between  these 
ratios,  would  probably  cool  the  gases  from  the  entering  tem- 
perature of  588  deg.  fahr.  to  the  requisite  240  deg.  fahr.  at 
exit. 

EXAMPLE. — The  Commonwealth  Edison  Co.,  Fisk  St.  Station  has  a 
nominal  boiler  horse  power  of  1,225  per  unit  and  an  economizer  surface 
of  8,500  sq.  ft.  per  unit,  or  a  ratio  of  (8,500  -5-  1,225)  =  approximately  7.1. 

308.  The  Rate  Of  Heat-Transfer  Between  The  Combustion- 
Gases  And  The  Water  In  An  Economizer  is  mainly  conditional 
upon  the  rate  of  the  gas-flow  through  the  economizer.     It  may 


270  STEAM  POWER  PLANT  AUXILIARIES  [Div.  8 

range  from  about  1.5  to  5.5  B.t.u.  per  hr.  per  sq.  ft.  of  heating 
surface  per  deg.  of  temperature-difference  between  the  gases 
and  the  water.  An  average  figure  for  the  rate  of  heat  transfer 
in  good  modern  economizer  installations  is  about  4.3  B.t.u. 
per  hr.  per  sq.  ft.  per  degree  difference  in  temperature.  It  is 
assumed  in  this  statement,  that  the  heating-surface  is  clean, 
and  that  there  is  no  air  infiltration  through  the  setting.  It  is 
also  presumed  that  the  flow-velocity  of  the  water  is  uniform 
for  various  rates  of  heat-transfer. 

NOTE.  —  The  minimum  rate  of  heat-transfer,  noted  above,  may  occur 
with  a  gas-flow  of  about  1,300  Ib.  per  hr.  per  sq.  ft.  of  heating-surface. 
The  maximum  rate  may  occur  with  a  unit  gas-flow  of  about  5,000  Ib. 
per  hr.  per  sq.  ft.  of  heating-surface.  Intermediate  rates  of  heat-transfer 
and  gas-flow  would  be  approximately  proportional  to  these  values. 

EXAMPLE  .  —  An  economizer  is  required  to  raise  the  temperature  of  75,000 
Ib.  of  water  per  hr.  through  15  deg.  fahr.  The  average  temperature- 
difference  between  the  combustion-gases  and  the  water  is  assumed  to  be 
200  deg.  fahr.  The  rate  of  heat-transfer  is  assumed  to  be  4  B.t.u.  per 
hr.  per  sq.  ft.  of  heating-surface  per  deg.  of  temperature-difference  be- 
tween the  gases  and  the  water.  The  assumed  specific  heat  of  the  water  = 
1.0.  What  is  the  requisite  area  of  heating-surface? 

SOLUTION.  —  The  hourly  transfer  of  heat  per  sq.  ft.  of  heating  surface  = 
(4  X  200)  =  800  B.t.u.  For  a  temperature-rise  of  15  deg.  fahr.,  each 
pound  of  water  must  absorb  (15  X  1.0)  =  15  B.t.u.  Therefore,  the 
requisite  heating-area  =  (75,000  X  15)  -f-  800  =  1,406  sq.  ft. 

309.  The  Percentage  Of  Fuel-Saving  Due  To  Economizer- 
Service  (Fig.  279)  may  be  computed  by  the  following  formula  : 


Wherein  X  =  per  cent,  of  saving.  T'/w  =*=  the  temperature 
of  the  feed-water  at  exit  from  the  economizer,  in  degrees 
Fahrenheit.  T/w  =  the  temperature  of  the  feed-water  at 
entrance  to  the  economizer,  in  degrees  Fahrenheit.  H  = 
the  total  heat  in  the  steam,  above  32  deg.  fahr.,  in  B.t.u.  per 
pound. 

EXAMPLE.  —  The  temperature  of  the  feed-water  entering  an  economizer 
is  160  deg.  fahr.  The  exit-temperature  of  the  water  is  305  deg.  fahr. 
The  boiler-pressure  is  200  Ib.  per  sq.  in.,  gage.  What  is  the  percentage 
of  fuel-saving,  due  to  the  economizer  service? 


SEC.  310] 


FUEL  ECONOMIZERS 


271 


SOLUTION. — The  total  heat  in  steam  at  200  Ib.  pressure  per  sq.  in., 
gage,  as  given  in  the  steam  tables  =  '1199.2  B.t.u.  per  Ib.  By  For.  (83): 
X  =  100(1%  -  T'fw)/[(H  +  32)  -  T'fw]  =  100  X  (305  -  160)  + 
[(1199.2  +  32)  -  160]  =  13.5  per  cent. 

NOTE. — It  is  commonly  assumed  that  an  economizer  will  effect  a 
fuel-saving  of  approximately  1  per  cent,  for  each  1 1-deg.  f ahr.  rise  in  the 
feed-water  temperature.  The  saving  may  amount  to  25  per  cent. 

EXAMPLE. — A  2,000-horse  power  boiler  plant  runs  10  hours  per  day  and 
300  days  per  year.  The  coal-consumption  is  4.6  Ib.  per  h.p.  per  hr.  The 
coal  costs  $4  per  ton.  It  is  assumed  that,  with  economizer-service,  13 
per  cent,  of  the  fuel  would  be  saved.  An  economizer-installation  would 


D   01    y 


o    f 


17  18  19  20  21  22  23  24  25  26  21  28  29  30  31 
Mont       h 


Fia.  279. — Diagram  Showing  Coal  Consumption  When  Economizer  Was  Stopped  And 
When  In  Operation  With  Clean  Tubes. 

cost  $9,500.  The  annual  cost  of  economizer-operation,  -depreciation 
and  -repairs  would  be  15  per  cent,  of  the  cost  of  installation.  What 
would  be  the  monetary  value  of  the  net  yearly  fuel-saving,  due  to  the 
economizer-service?  What  percentage  of  the  cost  of  the  economizer- 
installation  would  the  annual  saving  represent? 

SOLUTION.— The  annual  expense  for  fuel  =  [(2,000  X  4.6  X  10  X  300) 
-r-  2,000]  X  4  =  $55,200.  The  prospective  annual  cost  of  economizer 
operation,  depreciation  and  repairs  =  9,500  X  15  -5-  100  =  $1,425. 
Hence,  the  prospective  net  annual  saving  with  economizer-service  =  (55,200 
X  13  -=-  100)  -  1,425  =  $5,751.  This  represents  (5,751  -r-  9,500)  X 
100  =  60.5  per  cent,  of  the  cost  of  the  economizer-installation. 

310.  The  Increase  Of  Steam-Generating  Efficiency,  Due 
To  Economizer-Service  may  vary,  according  to  local  condi- 
tions, as  follows. 


272  STEAM  POWER  PLANT  AUXILIARIES  [Div.  8 

EXAMPLE. — When  the  boilers  are  run  slightly  above  their  rated 
capacity,  and  the  heating-surface  of  the  economizer  is  approximately 
equal  to  60  per  cent,  of  that  of  the  boilers,  the  efficiency-increase  may 
be  about  6  per  cent.  When  the  boilers  are  run  at  about  their  rated 
capacity,  with  the  combustion-gases  entering  the  economizer  at  a 
temperature  of  500  deg.  fahr.,  and  with  the  usual  flow-velocity,  the 
efficiency  increase  may  be  about  5  per  cent. 

EXAMPLE. — When  the  inherent  economy  of  the  boilers  is  good,  due  to 
excellence  of  design,  both  of  boilers  and  settings,  and  the  boilers  are  run 
at  200  per  cent,  of  their  rated  capacity,  the  efficiency-increase  may  be 
10  per  cent.  This  percentage  of  increase  is  based  on  a  temperature  of 
600  deg.  fahr.  and  a  flow-velocity  of  from  1,800  to  2,000  ft.  per  min.  for 
the  gases  entering  the  economizer.  Also,  the  economizer  heating-surface 
is  presumed  to  equal  70  per  cent,  of  the  boiler  heating-surface. 

EXAMPLE. — When  the  inherent  economy  of  the  boilers  is  poor,  due  to 
defective  design,  both  of  boilers  and  settings,  and  the  boilers  are  run  at 
less  than  their  rated  capacity,  the  efficiency-increase  may  be  nil;  there 
may,  under  such  conditions,  be  an  actual  decrease  in  overall  efficiency. 

311.  The  Principal  Advantages  Of  Economizer  Service  are : 
(1)  Fuel-saving.     The  saving  may  amount  to  from  5  to  over 
18  per  cent.     (2)  Increased  boiler-efficiency.     Where  the  boilers 
are  operated  at  over  200  per  cent,  of  their  nominal  rating,  and 
the  supply  of  exhaust  steam  for  heating  the  feed-water  is  scant, 
due  to  the  auxiliaries  being  electrically  driven,  the  increase  of 
efficiency  may  be  considerable.     (3)  Diminished  contraction 
stresses  in  steam  boilers.     This  results  from  the  high  feed-water 
temperature  which  is  attainable  with  the  economizer.     (4) 
Increased  flexibility  of  boiler  operation.     This  results  from  the 
storage-space  which  the  economizer  affords.     The  large  quan- 
tity of  hot  water  in  the  economizer  is  instantly  available  for 
use  in  the  event  of  a  sudden  overload. 

312.  The  Principal  Disadvantages  Of  Economizer  Service 
are:  (1)  Expense  for  installation.     (2)  Expense  for  maintenance. 
This  comprises,  mainly,  the  costs  of  repairing  and  operating 
(Sec.    294)    the    soot-scrapers    or    blowers.     (3)    Diminished 
overall  efficiency  of  the  plant.     This  may  occur  where  the  draft 
is  insufficient  for  economizer  operation,  or  where  the  boilers 
are  operated  below  their  nominal  ratings.     (4)  Bulkiness  of  the 
requisite  equipment.     An  economizer,  and  its  appurtenances, 
as  a  motor-  or  engine-driven  draft-fan,  requires  large  floor 
space.     If  the  economizer  is  erected  overhead,  much  altera- 


SEC.  313]  FUEL  ECONOMIZERS  273 

tion  of  piping  and  structural  details  will  generally  be  neces- 
sary to  make  room  for  the  equipment. 

313.  The  Conditions  Which  Usually  Determine  Whether  Or 
Not  Economizer  Should  Be  Installed  are  chiefly  as  follows : 

(1)  The  total  horse  power  of  the  plant  (Sec.  309). 

(2)  The  flue-gas  temperature.     Where  the  temperature  is  above  pos- 
sibly 450  to  550  deg.  fahr.,  profit  may  result  from  using  an  economizer. 
The  higher  the  temperature,  the  greater  the  saving.     The  flue  gases 
should  not  be  cooled  below  250  deg.  fahr.  because  the  vapor  in  the  gases 
will  be  condensed  on  the  economizer  tubes,  especially  near  the  exit. 
This  will  cause  soot  to  adhere  to  the  surfaces  of  the  tubes.     If  the  coal  is 
high  in  sulphur  content,  the  condensed  moisture  may  collect  sulphur 
dioxide  from  the  gases  and  dilute  sulphuric  acid  result.     This  corrodes 
the  tubes  (Sec.  292). 

(3)  The  boiler  pressure.     When  the  pressure  is  250  Ib.  per  sq.  in.,   or 
over,  an  economizer  is  practically  indispensable. 

(4)  The  character  of  the  load.     If  the  plant  is  heavily  overloaded,  either 
steadily  or  intermittently,  there  may  be  need  for  an  economizer.     Gener- 
ally a  substantial  saving  may  be  realized  when  boilers  are  operated  well 
above  their  ratings  for  a  large  proportion  of  the  time. 

(5)  The  feed-water  temperature.     Increased  economy  may  result  from 
the  high  temperature  which  may  be  obtained  with  an  economizer.     A 
great  saving  should    result  in  plants  running  condensing  when  motor- 
driven  auxiliaries  are  employed.     Under  these  conditions  there  is  usu- 
ally insufficient  exhaust  steam  to  heat  the  feed-water.     The  economizer 
should  deliver  the  water  at  a  much  higher  temperature — much  higher 
than  210  deg.  fahr.  which  is  usually  the  limit  for  exhaust  steam  heaters. 
The  wider  the  range  over  which  the  economizer  heats  the  water,  usually 
the  greater  is  the  saving. 

(6)  The  quantity  of  exhaust-steam   available  for  heating.     When  feed- 
water  heaters  are  installed  and  plenty  of  exhaust  steam  is  available, 
which  would  be  lost  if  not  used  in  the  heater,  an  economizer  may  not 
show  any  considerable  saving. 

(7)  The  quality  of  the  feed  water.     If  the  water  contains  impurities 
which  will  form  scale  in  an  economizer,  then  an  economizer  may  prove 
undesirable. 

(8)  The  available  means  for  supplying  sufficient  draft,   and  the  cost 
thereof.     Lack  of  building  space   might  necessitate  erection  of  a  tall 
chimney.     Otherwise,  an  artificial  draft  equipment  may  be  necessary. 
The  cost  of  a  tall  chimney  might  be  prohibitive.     Likewise,  the  expendi- 
ture of  from  1  to  4  per  cent,  of  the  total  power  output  for  driving  the 
draft  equipment  (see  author's  STEAM  BOILERS)  might  be  prohibitive. 

(9)  The  initial  cost  of  the  economizer.     When  economizers  are  required 
to  sustain  pressures  greater  than  250  Ib.  per  sq.  in.,  their  cost  increases 
rapidly  with  the  pressure. 

18 


274  STEAM  POWER  PLANT  AUXILIARIES  [Div.  8 

(10)  The  price  of  coal.  When  the  price  of  coal  is  high  there  is  more 
saving  than  when  it  is  cheap,  unless  the  cost  of  the  economizer  and  its 
operation  also  are  high  in  the  same  proportion. 

314.  Economizers     Should    Be     Inspected     Periodically. 

There  should  be  a  monthly  overall  detail  inspection.  Certain 
elements  may  require  more  frequent  attention.  Inspection 
should  cover  the  following  details : 

(1)  The  external  surfaces.     Leaks  and  soot-deposits  should  be  looked 
for.     Soot  should  be  removed.     Leaks  should  be  stopped.     They  cause 
corrosion  and  tend  to  produce  soot-  and  rust-scale  on  the  tubes. 

(2)  The  internal  surfaces.     A  scale-forming  tendency   (Sec.   295)   in 
the  tubes  should  be  looked  for.     If  such  exists,  the  economizer  should  be 
opened  as  frequently  as  practicable  and  the  tubes  washed  with  a  hose. 

(3)  The  safety-valve.     Corrosion  between  the  valve  and  seat  should  be 
looked  for.     Also,  the  valve  mechanism  should  be  tested  for  freedom  of 
movement. 

(4)  The  blow-off  valves.     The  valve  discs  should  be  examined.     Like- 
wise the  packing  of  the  stems.     If  the  packing  is  dry  and  hard,  the  stems 
should  be  repacked.     The  stems  should  work  freely. 

(5)  The  flange-joints.     Leaking  joints  should  be  repacked.     Also,  any 
straining  effect  on  the  joints,  due  to  restraint  of  expansion  and  contrac- 
tion in  the  pipe-lines,  should  be  rectified. 

(6)  The  soot-scrapers  or  blowers.     The  distance  traveled  by  the  scrapers 
should  be  noted.     It  should  be  the  full  length  of  the  tubes.     The  blower- 
nozzles  should  be  examined.     If  the  nozzle-orifices  have  been  enlarged 
by  erosion,  the  nozzles  should  be  renewed. 

(7)  The  gearing  and  reversing-mechanism  of  the  soot-scrapers.     Broken 
gear-teeth  should  be  looked  for.     The  security  of  bolts,  pins  and  cotters 
should  be  tested.     The  lubrication  of  the  bearings  should  be  noted. 

(8)  The   dampers.     The   devices   for   damper-adjustment   should   be 
tested. 

(9)  The  setting.     Search  should  be  made  for  cracks  in  the  masonry. 
Leakage  of  air  around  door-frames  should  be  looked  for.     Suspected 
places  may  be  tested  by  applying  the  flame  of  a  candle  or  torch,  with  the 
stack  damper  open. 

(10)  The  soot-pits  or  chambers.     These  should  be  entirely  emptied  of 
their  contents  as  frequently  as  is  necessary. 

(11)  The  thermometers.     The  accuracy  of  the  instruments  should  be 
noted. 

NOTE. — EXAMINATIONS  FOR  EVIDENCES  OF  PITTING  ON  THE  INTERIOR 
SURFACES  OF  ECONOMIZERS  (see  the  author's  STEAM  BOILERS)  should  be 
made  annually.  Facility  in  making  these  inspections  may  sometimes 
necessitate  a  partial  disassembling  of  the  economizer  sections. 


SEC.  315]  FUEL  ECONOMIZERS  275 

315.  The  Cost  Of  An  Economizer  And  Its  Installation  is 
usually  computed  on  the  ratio  of  the  economizer  heating- 
surface  to  the  boiler  horse  power.  When  this  ratio  is  5:1  the 
cost,  prior  to  the  Great  War,  was  about  $6  per  boiler  horse 
power.  When  the  ratio  is  4.8:1,  the  prewar  cost,  for  plants 
containing  1000  boiler  horse  power,  or  more,  was  about  $5.50 
per  boiler  horse  power.  Otherwise,  the  cost,  regardless  of 
either  the  size  of  the  installation  or  the  ratio  mentioned  above, 
may  be  based  directly  on  the  extent  of  economizer  heating- 
surface.  On  this  basis,  a  prewar  cost  of  $1.20  per  sq.  ft.  was 
usually  assumed.  POWER  PLANT  ENGINEERING,  Apr.  1,  1920, 
states  that  cost,  including  fan,  motors,  etc.,  will  now  average 
about  $4  per  sq.  ft.  of  economizer  surface. 

QUESTIONS  ON  DIVISION  8 

1.  What  is  the  function  of  a  fuel-economizer? 

2.  Describe  an  independent  fuel-economizer.     An  integral  fuel-economizer. 

3.  What   materials  are  used  in  economizer   construction?     What  are  the  relative 
advantages  and  disadvantages  of  these  materials? 

4.  What  detrimental  effects  may  result  from  coatings  of  soot  on  economizer  surfaces? 

5.  In  what  manner  do  sedimental  deposits  in  economizer-tubes  affect  the  economy 
of  the  apparatus? 

6.  What  physical  injury  may  result  from  impurities  in  the  water  pumped  through 
an  economizer? 

7.  Describe  the  operation  of  economizer  tube  scrapers. 

8.  How  may  mineral  substances  in  the  feed-water  be  prevented  from  forming  scale 
in  economizer-tubes? 

9.  Why  does  not  hard  scale  form  as  readily  in  an  economizer  as  in  a  boiler? 

10.  Explain  the  ill-effects  of  air-infiltration  through  an  economizer-setting. 

11.  How  may  air-infiltration  through  an  economizer-setting  be  detected? 

12.  How  does  air-infiltration  affect  the  quality  of  the  combustion-gases  in  an  econo- 
mizer?    What  may  be  regarded  as  a  reasonable  drop  in  the  percentage  of  COz  in  the 
gases? 

13.  If  a  boiler  plant  is  being  operated  with  natural  draft,  what  would  be  the  probable 
effect  on  the  draft  if  an  economizer  were  installed? 

14.  What  is  the  usual  method  of  supplying  draft  for  boiler  plants  which  are  equipped 
with  economizers? 

15.  Enumerate    the    principal    advantages    of    economizer    service.     The    principal 
disadvantages. 

16.  Enumerate  the  conditions  of  boiler-service  which  chiefly  determine  the  advis- 
ability of  installing  an  economizer. 

17.  Enumerate  the  structural  details  and  service-conditions  toward  which  economiz- 
er-inspections should  be  particularly  directed. 

18.  Explain  why  economizer  heating-surface  is  more  effective  in  absorbing  heat  from 
the  cornbustion-gases  leaving  a  boiler  than  an  extension  of  boiler  heating-surface  would 
be. 

19.  What  is  a  contra  flow  in  an  economizer  installation?     A  parallel  flow?     Which  is 
the  more  effective? 

20.  What  is  the  average  ratio,  in  economizer  operation  in  general,  of  the  drop  of 
combustion-gas  temperature  to  the  rise  of  feed-water  temperature?     Give  some  ex- 
amples,   approximating    this   ratio,    as    observed    in   typical   economizer-installations. 


276  STEAM  POWER  PLANT  AUXILIARIES  [Div.  8 

21.  What   service-condition   mainly   determines   the   extent   of   economizer   heating- 
surface  that  can  be  profitably  used? 

22.  Upon  what  service-condition  is  the  rate  of  heat-transfer  in  an  economizer  mainly 
contingent? 

23.  What  percentages  of  increase  of  steam-making  efficiency,  for  various  conditions  of 
boiler-service,  may  ordinarily  be  anticipated  from  economizer-service? 

PROBLEMS  ON  DIVISIONS 

1.  The  combustion-gases  leaving  a  boiler  have  a  temperature  of  550  deg.  fahr.  and  a 
COz  content  of  12  per  cent.     On  account  of  leakage  of  air  through  the  setting,  the  gases 
leaving  the  economizer  have  a  temperature  of  250  deg.  fahr.  and  a  COi  content  of  8  per 
cent.     The  specific  heat  of  the  gases  =  0.24.     What  percentage  of  the  heat  is  lost  when 
the  temperature  of  the  outside  air  is  50  deg.  fahr.? 

2.  For  each  15  Ib.  of  combustion-gases  flowing  through  an  economizer  there  is  a 
corresponding  water-flow  of  8  Ib.     What  should  be  the  ratio  of  the  decrease  of  combus- 
tion-gas temperature  to  the  increase  of  feed-water  temperature? 

3.  A  boiler  is  to  deliver  steam  at  a  pressure  of  175  Ib.  per  sq.  in.,  gage.     The  tempera- 
ture of  steam  at  this  pressure  =  377.5  deg.  fahr.     The  boiler  will  be  run  24  hr.  per  day. 
Coal  will  cost  $3.00  per  ton.     How     much  heating-surface,  per    boiler    horsepower, 
should  the  boiler  have  in  order  that  it  may  be  operated  economically? 

4.  It  is  assumed  that  an  economizer  is  to  be  installed  in  connection  with  the  boiler 
mentioned  in  Problem  3.     It  is  also  assumed  that  the  feed-water  will  enter  the  economiz- 
er at  a  temperature  of  200  deg.  fahr.     What  should  be  the  least  temperature,  consistent 
with  economical  operation,  of  the  combustion  gases  at  exit  from  the  economizer? 

5.  The  average  temperature-difference  between  the  feed-water  and  the  combustion- 
gases  in  an  economizer  is  assumed  to  be  300  deg.  fahr.      The  economizer  is  required  to 
raise  the  temperature  of  50,000  Ib.  of  water  per  hour  through  50  deg.  fahr.     The  rate  of 
heat  transfer  is  assumed  to  be  5.5  B.t.u.  per  hr.  per  sq.  ft.  of-heating  surface  per  degree 
of  temperature-difference  between  the  gases  and  the  water.     The  assumed  specific  heat 
of  the  water  =  1.0.     What  should  be  the  area  of  the  economizer  heating-surface? 

6.  The  temperature  of  the  water  entering  an  economizer  is   110  deg.   fahr.     The 
temperature  of  the  water  at  exit  is  250  deg.  fahr.     The  boiler-pressure  is  175  Ib.  per  sq. 
in.,  gage.     The  total  heat,  above  32  deg.  fahr.,  in  steam  at  this  pressure  =  1197.3  B.t.u. 
per  Ib.    What  is  the  percentage  of  fuel-saving? 

7.  A  2400-horse  power  boiler  plant  runs  24  hr.  per  day  and  300  days  per  year.     The 
coal-consumption  is  4.3  Ib.  per  h.p.  per  hr.     The  coal  costs  $4.25  per  ton.     It  would 
cost  $12,000  to  install  an  economizer  in  this  plant.     The  annual  cost  of  operation, 
maintenance  and  depreciation  would  amount  to  15  per  cent,  of  the  cost  of  installation. 
Assuming  that  the  economizer  would  effect  a  fuel-saving  of  12.3  per  cent.,  what  would 
be  the  monetary  value  of  the  saving  per  year? 


DIVISION  9 


STEAM  CONDENSERS 

316.  A  Steam  Condenser  As  Used  In  Connection  With  A 
Steam  Engine  is  a  device  for  reducing  exhaust  steam  to  water. 
The  purpose  of  a  condenser  is  to  increase  the  power  which  is 
developed  by  an  engine  from  a  given  quantity  of  steam;  or, 
conversely,   to  decrease  its  steam  consumption  for  a  given 
power  output. 

NOTE. — A  CONDENSER  INCREASES  ENGINE  ECONOMY  BY  CREATING  A 
PARTIAL  VACUUM  in  a  container  into  which  the  engine  discharges  its 
exhaust  steam.  The  method  of  creating  the  vacuum  is  to  cool  the  ex- 
haust steam  sufficiently  so  that 
it  will  condense  to  water,  which 
occupies  very  much  less  space. 
The  effect  of  the  partial  vacuum 
thus  created  is  to  give  the  engine 
10  to  15  Ib.  per  sq.  in.  more  work- 
ing pressure  without  any  in- 
crease in  boiler  pressure  or 
material  increase  in  fuel  con- 
sumed. Greater  working  pres- 
sure with  a  given  quantity  of 
steam  results  in  greater  power 
output. 

317.  How  A  Condenser 
Increases     The     Working 
Pressure  of  a  steam   cyl- 
inder may  be  demonstrated 
thus:— Fig.  280  shows  two 
elementary  steam  cylinders 
surrounded  by  air  at  nor- 
mal atmospheric  pressure.    This  pressure  is  equal  to  about  14.7 
Ib.  per  sq.  in.  at  sea  level.     That  is,  any  object  exposed  to  the 
air  at  sea  level  has  14.7-lb.-per-sq.-in.  pressure  exerted  on  it 
from  all  directions.     Assume  that  a  pressure  of  100  Ib.  per  sq. 

277 


.-Stecrm  Supply lOOLb..  . 
Per  5cf.ln(Gcrgre)-  •  • 


FIG.  280. — Showing  The  Effect  Of  Vacuum 
Produced  By  Condensation  On  The  Working 
Pressure  Of  A  Steam  Piston  Having  An  Area 
Of  1  Sq.  In. 


278 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


in.  as  indicated  by  a  steam  gage,  is  exerted  on  the  under  side  of 
a  piston,  A,  of  one  square  inch  area.  Then  the  piston  will 
have  a  lifting  force  of  100  Ib.  There  is  really  114.7  Ib.  pressure 
pushing  on  the  under  side  and  14.7  Ib.  on  the  upper,  but  only 
the  difference,  100  Ib.  is  effective.  .  But,  suppose  the  space 
above  piston,  B,  which  has  the  same  area,  is  first  filled  with 
steam  and  then  condensed  (assuming  that  it  will  condense 
completely).  Then  there  will  be  no  pressure  on  the  upper 
side  of  the  piston  and  the  whole  114.7  Ib.  on  the  lower  side 
becomes  effective.  The  piston  then  has  a  lifting  force  of 
114.7  Ib.  Thus,  by  condensing  the  steam  above  the  piston  in 
B  its  lifting  force  is  increased  by  14.7  Ib. 

318.  Power  Was  Developed  By  Condensation  in  Primitive 
Steam  Engines. — The  primitive  engine  of  Newcomen  shown  in 

Fig.  281  works  entirely  by  con- 
densation. 

EXPLANATION. — On  the  upstroke 
steam,  at  a  little  above  atmospheric 
pressure,  flows  into  the  cylinder,  C, 
through  the  valve,  7,  and  is  then 
condensed  by  a  jet  of  cold  water,  S, 
admitted  at  V.  The  partial  vacuum 
thus  created  allows  the  atmospheric 
pressure  above  to  force  the  piston, 
P,  down  so  that  power  is  developed 
on  the  downstroke.  The  condensed 
steam  and  condensing-water  are 
drained  out  through  valve,  0,  while 
the  piston  is  on  the  upstroke.  A 
weight,  W,  counterbalances  the  piston 
and  is  connected  to  a  pump  rod,  77, 
which  does  the  work  of  the  engine.  By  alternately  admitting  steam 
into  C  and  condensing  it  therein,  rod,  77,  is  forced  to  move  up  and  down 
and  thereby  pump  water. 

319.  An  Improvement  Of  Newcomen's  Engine  Was  Made 
By  Watt  (Fig.  282)  who  condensed  the  steam  in  a  separate 
chamber  or  condenser,  D,  with  a  jet  of  cold  water.     The  chief 
advantage  of  this  arrangement  over  the  former  is  that  in  this 
the  cylinder  remains  hot  and  the  condenser  cold  so  that  neither 
has  to  be  alternately  heated  and  cooled,  as  with  Newcomen's 
arrangement.     Watt    then    made    his    engine    double-acting 


FIG.  281. — Newcomen's   Condensation 
Engine.      (Year  1763.) 


SEC.  320] 


STEAM  CONDENSERS 


279 


(Fig.  283)  and  operated  the  valves,  /,  J,  0  and  Q,  by  a 
system  of  levers.  The  condenser,  D,  then  acted  continuously 
and  the  condensed  water  had  to  be  pumped  out  of  the  conden- 
ser. Small  amounts  of  air  also  accumulated  in  the  condenser 
and  were  pumped  out. 


Condensing 
Wafer-. 


FIG.  282. — Watt's  Condensation  Engine. 


FIG.    283. — Watt's  Double-Acting  Con- 
densing Engine. 


320.  How  The  Condenser  Saves  Steam  may  be  shown  by 
either  of  two  methods:  (1)  By  comparing  the  thermal  effi- 
ciency of  condensing  and  non-condensing  operation  (Sec.  321). 
(2)  By  computing  the  ratio  of  mean  effective  pressures  for  con- 
densing and  non-condensing  operation  (Sec.  322).  The  first 
method  is  based  on  the  work  done  by  the  steam  in  expanding 
and  is  therefore,  not  accurately  applicable  to  those  slide-valve 
and  similar  engines  in  which  the  expansion  is  either  zero  or 
small.  The  ideal  conditions  which  this  method  assumes  are 
approached  in  the  compound  Corliss  engine  and  the  steam 
turbine.  The  second  method  gives  a  fair  estimate  of  the  actual 
power  increase  effected  by  the  condenser  except  that  the  power 
required  by  the  condenser  auxiliaries  must  be  deducted. 

NOTE. — The  relation  of  the  power  developed  by  a  condensing  engine 
to  that  by  a  corresponding  non-condensing  engine  is  shown  in  Fig.  284. 
The  area  DCFE  represents  the  increase  in  power  due  to  the  condenser. 
This  may  be  compared  to  the  power  developed  by  the  primitive  con- 
densation engine  (Fig.  281).  The  area  GABCD  represents  the  power 
developed  by  the  engine  when  operating  non-condensing.  The  modern 


280 


STEAM  POWER  PLANT.  AUXILIARIES 


[Div.  9 


condensing  engine  develops  the  sum  of  these  amounts  of  power  from  the 
same  quantity  of  steam,  as  represented  by  the  total  area  GABEF. 


{-Area  Of  Non-Condensing  Card=ABCDG  = 


Work  Done  During  Non-Condensing  Stroke 
'  \-Area  Of  Condensing  Car d=  ABE FG=  Work 


Done  During  Conc/enstng  Stroke    ,.,.., 
Non -Condensing  Back  Pressure  Li ne= 


Condensing  Back- Pressure  Lme-Ef. 
Saying  In  Back  Pressure=FD=12Lb./5aJ, 


Area  DCE f=  Gain  In  Work  Done,  Due  T&t-r 
Condenser. 


•Mean-EfrecTwe  Pressure 
Non-Condensi, 


•Atmospheric  Pressure 


Length  Of  Stroke 

FIG.  284. — Theoretical  Indicator  Cards  Showing  Work  Gained  By  Condensing  Over 
Non-Condensing  Operation.  (Gain  in  work  of  condensing  over  non-condensing 
operation  =33  per  cent.) 


!x» 


50 


ing  Back 

fm 


rrww  Of  'Non-Condensing  Card±  GABCDG_ 
Work  Done  During  Non-Condens/ng  Stroked 
tfArea  Of  Condensing  Card=  GABCDG  =m>n 


e  During  Condensing  Stroke^ 


Operation  . .  .^  ......   ....  --.-.-- 

Virh  Non-Condensing  At  A.  The  Additional 
Steam  Required  As  Represented  By  A  A'  - 
Indicates  The  Greater  Steam  Consumption 
Of  Non-Condensing  Over  Condensing. 


k  Pressure  JL/ne 


\Condensinqf,  Back  Pressure  ///7g— ^ 


K Length  Of  Stroke •• -H 

Fia.  285. — Theoretical  Indicator  Cards  Showing  Difference  In  Steam  Consumption  Of 
Condensing  And  Non-Condensing  Operation  For  Equal  Work  Area. 

When  steam  is  used  expansively  (as  indicated  by  the  curved  expansion 
line  A 5)  a  given  difference  in  pressure  below  the  atmospheric  line 
(such  as  P3,  Fig.  284)  represents  much  more  power  than  a  similar  differ- 


SEC.  321]  STEAM  CONDENSERS  281 

ence  in  pressure  above  the  boiler  pressure,  P4.  In  other  words,  13  Ib.  per 
sq.in.  of  vacuum  in  a  condenser  increases  the  power  of  a  good  engine  much 
more  than  13  Ib.  per  sq.  in.  more  boiler  pressure. 

NOTE. — The  areas  in  the  indicator  cards  (Figs.  284  and  285)  represent 
energy  or  work  delivered  during  one  stroke  of  an  engine  but,  assuming 
a  constant  engine  speed,  they  also  represent  the  proportional  power 
developed. 

321.  The  Increase  In  Thermal  Efficiency  Effected  By  A 
Condenser   may   be   estimated   by   the   steam   temperature 
relations.     The  greatest  possible  thermal  efficiency  of  any 
heat  engine  is  represented  by  the  equation : 

(84)  E,  =  ^  — 2  (decimal) 

Tj 

The  efficiency  which  this  formula  gives  may  be  approximated 
by  an  actual  engine  but  can  never  be  attained.  It  might  be 
attained  only  by  an  ideally-perfect  engine.  (See  the  author's 
PRACTICAL  HEAT)  Wherein:  Et  =  the  greatest  possible  ther- 
mal efficiency  of  any  heat  engine.  TI  =  the  absolute  tempera- 
ture at  which  the  steam  is  admitted.  T2  =  the  absolute 
temperature  at  which  the  steam  is  exhausted.  -Absolute 
temperature  =  460  deg.  +  the  temperature  in  deg.  fahr. 

EXAMPLE. — Assume  an  engine  using  saturated  steam  at  115  Ib.  per 
sq.  in.  abs.  As  shown  by  a  steam  table  this  steam  has  a  temperature  of 
338  deg.  fahr.  or  338  +  460  =  798  deg.  fahr.  abs.  It  exhausts,  when 
running  non-condensing,  at  a  pressure  of  16  Ib.  per  sq.  in.  abs.;  and, 
when  running  condensing  at  2  Ib.  per  sq.  in.  abs.  These  pressures,  as 
shown  by  a  steam  table,  correspond  to  676  and  586  deg.  abs.  respectively. 
Hence,  by  For.  (84),  the  ideal  efficiency  non-condensing  is:  E<  =  (Tx  — 
Tj)/Ti  =  (798  -  676) /798  =  15.3  per  cent.  The  ideal  efficiency  con- 
densing =  Et  =  (798  -  586) /798  =  26.6  per  cent. 

322.  The  Theoretical  Saving  In  Power  Due  To  The  Use  Of 
A.  Condenser  may  be  computed  by  the  following   formula: 

4Q  Pi 

(85)  Saving  =  -  -^  (per  cent.) 

•Lm 

Wherein:  Phmv  =  the  vacuum  obtained  in  the  condenser,  in 
inches  of  mercury.  Pm  =  the  mean  effective  pressure  of  the 
engine  running  non-condensing,  in  Ib.  per  sq.  in. 

NOTE. — The  saving  is  much  more  than  proportional  to  the  increase  in 
working  pressure  of  the  engine.  That  is  (Fig.  284)  the  saving  is  propor- 
tional to  P3  -s-  PZ  not  to  PS  •£•  Pi.  The  mean  effective  pressure  is  found 


282 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


in  practice  by  measuring  the  area  of  an  actual  indicator  diagram  with  a 
planimeter  and  dividing  this  area  by  the  length  of  the  diagram.  If  this 
value  (represented  by  P2,  Fig.  284)  is  multiplied  by  the  constant  of  the 
spring  used,  (for  instance  80  for  an  80-lb.  spring)  the  mean  effective  pres- 
sure in  pounds  per  square  inch,  POT,  For.  (85),  will  be  obtained.  The 
pressure  difference  due  to  the  condenser  (P3,  Fig.  284)  applies  evenly 
throughout  the  stroke  and  so  the  vacuum  obtained  in  the  condenser 
may  be  taken  as  proportional  to  P3  if  reduced  also  to  pounds  per  square 
inch.  The  vacuum  expressed  in  inches  of  mercury  may  be  reduced  to 
pounds  per  square  inch  by  dividing  by  2.03.  The  saving,  then  = 

PA«*  49P*™ 


(86) 


100  X 


X2.03-       Pm 

Example. — The  boiler  pres- 
sure of  an  engine  is  100  Ib.  per 
sq.  in.  gage  and  the  mean  effec- 
tive pressure  of  the  engine  run- 
ning non-condensing  is  44.6  Ib. 
per  sq.  in.  What  theoretical  sav- 
ing will  result  from  condensing 
operation  in  a  26  in.  vacuum? 

SOLUTION. — By  For.  (85),  the 
saving  =  49  Phmv/Pm  =  49  X 
26/44.6,  28.6  per  cent. 

NOTE. — Fig.  285  shows  the 
effect  of  the  condenser  on  en- 
gine economy  keeping  the 
amount  of  power  developed 
constant,  instead  of  keeping 
the  amount  of  steam  used  con- 
stant as  on  Fig.  284. 


323.  The  Steam  Saving 
Due  To  A  Condenser  On 
The  Basis  Of  Decreased 


20     40     60     80     100     120    WO    160    180  200 
Initial   Steam   Pressure  Absolute 
UxPer   Sq.  In 

FIG.    286.— Diagram    Showing    The    Steam  Back   PrCSSUre  may  be  de- 
Consumption  Of  A  Perfect  Steam  Engine  When  ^ermined     VCFV     closely     by 
Receiving  Steam  At  Different  Pressures  And  , 
Exhausting  Against  Different  Back  Pressures,  the  USC   OI    a  Suitable  graph 
(Note  bad  effect  of  high  back  pressure.     The  (Y[g.   286) .       First  the  IlOn- 
steam  consumption  of  actual  engines  is  affected 
in    about    the    same    proportion.)     (Harrison  Condensing  Steam  COnSUmp- 

safety  Boiler  Works  Catalog.)  tjon  js  determined  from  the 

graph,  which  gives  values  for  an  ideal  engine.  Similarly  the 
ideal  condensing  consumption  is  determined.  Then  the  ratio 
of  these  values  is  applied  to  the  actual  steam  consumption 
considered. 


SEC.  324] 


STEAM  CONDENSERS 


283 


EXAMPLE. — Consider  an  engine  working  at  100  Ib.  per  sq.  in.  abs.  with 
1  Ib.  per  sq.  in.  gage  back  pressure,  consuming  25  Ib.  of  steam  per  h.p. 
hr.  How  much  steam  will  it  use  if  operated  condensing  with  a  26  in. 
vacuum? 

SOLUTION. — Locate  ordinate  A  (Fig.  286)  corresponding  to  100  Ib.  per 
sq.  in.  abs.  on  the  lower  scale  and  on  this  ordinate,  find  B  and  C  corre- 
sponding to  1  Ib.  gage  per  sq.  in.  and  26  in.  of  vacuum.  The  correspond- 
ing ideal  consumptions  are  about  19  and  10  Ib.  per  h.p.  hr.  That  is,  the 
ideal  condensing  consumption  is  10/19  of  the  ideal  non-condensing  con- 
sumption. But  the  engine  actually  consumes  25  Ib.  per  h.p.  hr.  when 
discharging  against  a  1  Ib.  per  sq.  in.  back  pressure.  Hence,  the  actual 
consumption  at  26  in.  vacuum  =  25  X  10/19  =  13.2  Ib.  per  h.p.  hr. 

324.  The  Function  Of  A  Condenser  Air-Pump  (P,  Fig.  287) 
is  to  produce  and  maintain  a  vacuum  in  the  condenser  by 
removing  the  air  which  enters  with  the  exhaust  steam.  The 
air  may  leak  into  the  exhaust  system  in  various  ways,  as 
through  imperfectly-packed  pipe- joints  and  stuffing-boxes. 
Also,  it  may  pass  into  the 
boilers  with  the  feedwater, 
and  thence  become  entrained 
with  the  steam-supply  to  the 
engine.  Air  if  permitted  to 
collect  in  the  condenser 
would  obviously  prevent  the 
production  of  an  effective 
vacuum. 


•2.7%" Absolute 
l2*/o"Pressare.  Difference-^ 


1.2  Cu.Ft  OfAfc 
A'irPump — -> 
Relief  \bdv 


lib.  Or  155  Cu.  Ft.  Of  Steam-' 
l2°/a"PreS5ure  Difference. 


I26Lb.OrAboufZ 
Cu.rt.0f  Water 

drcu/Cftinp 

FIG.  287. — Diagram  Of  Elementary  Jet 
Condenser  Showing  Relative  Volumes  And 
Pressures  Of  Air,  Water  And  Steam.  (On 
the  basis  of  one  pound  of  steam.) 


EXPLANATION.  —  Imagine  a 
closed  vessel  to  contain  a  perfect 
vacuum,  and  that  a  quantity  of 
steam,  unmixed  with  air  or 
non-condensible  vapors,  be  ad- 
mitted thereto.  Then,  if  the  steam  be  cooled  to  a  temperature  of 
110  deg.  fahr.,  the  'absolute  pressure  in  the  condenser  (Table  345) 
due  to  the  presence  of  the  still  uncondensed  water  vapor  would  be  2.6 
in.  of  mercury.  Or,  referred  to  a  30-in.  barometer  a  partial  vacuum  of 
30.0  —  2.6  =  27.4  in.  of  mercury  would  result.  But  if  air  had  been 
mixed  with  the  steam  that  was  admitted  to  the  vessel,  then  the  air  of 
itself  would  exert  a  pressure  in  addition  to  that  due  to  the  water 
vapor  and  would  decrease  the  vacuum  which  would  otherwise  obtain. 
Hence,  with  air  in  the  condenser,  a  partial  vacuum  of  27.4  in.  of  mercury 
could  not  result  from  condensation  of  the  steam.  The  degree  of  vacuum 
actually  obtainable  would  depend  upon  the  quantity  of  air  present. 


284 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


325.  The  Power  Required  To  Remove  The  Air  And  Water 
From  A  Condenser  is,  as  will  be  shown,  relatively  small  (Fig. 
287)  compared  to  the  power  developed  by  the  condenser: 
First  estimate  the  power  developed  by  1  Ib.  of  steam  in  a 
condenser  under  typical  conditions.  One  pound  of  steam  at 
an  absolute  pressure  of  2.7  Ib.  per  sq.  in.  occupies  about  135 
cu.  ft.  The  theoretical  work  done  by  the  engine  due  to 
condensation  with  a  vacuum  of  12  Ib.  per  sq.  in.,  which  cor- 
responds to  about  2.7  Ib.  per  sq.  in.  abs.  then,  =  135  X  144 
X  12  =  233,000  ft.  Ibs.  for  each  pound  of 
steam  condensed.  If  126  Ib.  (a  rather-high 
value)  of  water  are  required  to  condense 
the  1  Ib.  of  steam,  the  volume  of  water  to 
be  pumped  out  is,  since  there  are  63  Ib.  of 
water  in  1  cu.  ft.,  2  cu.  ft.  The  theoretical 
work  done  by  pump  C  in  pumping  out  the 
water  therefore,  =  2  X  144  X  12  =  3,450 
ft.  Ib.  for  each  pound  of  steam.  If  the 
volume  of  the  air  at  condenser  pressure  is 
60  per  cent,  of  that  of  the  water,  the  work 
required  to  remove  it,  =  1.2  X  144  X  12  = 
2,070  ft.  Ib.  for  each  pound  of  steam.  The 
theoretical  power  required  to  remove 
the  air  and  water,  then,  =  100  (3,450  -f- 
2,070)  /233,000  =  2.4  per  cent.  In  large 
plants  the  actual  steam  required  to  drive 
the  condenser  auxiliaries,  when  they  are 
steam  driven,  amounts  to  about  1  to  3  per 
cent,  of  that  required  by  the  main  engine. 
326'  Gages  for  Measuring  Condenser 


10- 


FIG.  288.-Mercury 


Vacuum  Gage  which    Vacuum    are   of    two   principal  types:  (1) 


l^d*  i°n  Bourdon  tube  vacuum  gages.  (2)  Mercury 
Pounds  Per  Square  vacuum  gages,  or  manometers.  Both  types 
usually  read  in  inches  of  mercury.  The 
principle  of  the  mercury  vacuum  gage  (Fig.  288)  is  similar 
to  that  of  the  barometer  (See  the  author's  PRACTICAL  HEAT). 
The  barometer  has  practically  zero  pressure  above  the  mer- 
cury while  with  the  vacuum  gage  the  pressure  to  be  measured 
is  above  the  mercury. 


SEC.  327]  STEAM  CONDENSERS  285 

NOTE. — Vacuum  gages  of  both  the  Bourdon-tube  and  the  mercury 
types  indicate  the  difference  in  pressure  between  the  condenser  and  the 
outside  air;  and  not  the  absolute  pressure.  Therefore  the  absolute  pres- 
sure will  be  different  for  a  given  vacuum  gage  reading  under  different 
weather  conditions  and  at  different  altitudes.  When  the  barometer  is 
low,  a  condenser  will,  for  given  cooling-water  supply,  efficiency  and  other 
conditions;  give  less  vacuum  (but  the  same  absolute  pressure)  than  when 
the  barometric  pressure  is  high. 

327.  The   Absolute   Pressure   In  A   Condenser  May  Be 
Computed  From  The  Reading   Of  The  Vacuum  Gage  by 

applying  the  following  formula : 

(87)  Pa  =  Phmh^hmv         (pounds  per  sq.  in.) 

Wherein:  Pa  =  the  absolute  condenser  pressure,  in  pounds 
per  square  inch.  Phmb  =  the  barometer  reading,  in  inches  of 
mercury.  Phmv  =  the  vacuum-gage  reading,  in  inches  of 
mercury.  2.03  =  the  height,  in  inches,  of  a  mercury  column 
which  exerts  a  pressure  of  1  Ib.  per  sq.  in. 

NOTE. — If  the  barometer  reading  is  corrected  for  temperature,  the 
vacuum  gage  reading  should  be  corrected  for  temperature  also.  If  both 
vacuum  gage  and  barometer  use  mercury  columns  referred  to  brass  scales 
the  error  due  to  neglecting  temperature  in  this  formula  will  not  be  appre- 
ciable. 

EXAMPLE. — A  condenser  vacuum-gage  reads  26  in.  while  the  barometer 
reads  29.4  in.  What  is  the  absolute  condenser  pressure?  What  is  the 
degree  of  vacuum,  as  a  per  cent,  of  that  which  is  theoretically  possible? 

SOLUTION.— By  For.  (87),  P«  =  (Phmb  -  Phmv)/2.03  =  (29.4  -  26)  -=- 
2.03  =  1.67  Ih.  per  sq.  in.  The  degree  of  vacuum,  as  referred  to  that  which 
is  theoretically  possible  in  this  case  =  (26  -f-  29.4)  X  100  =  88.4  per  cent. 

328.  The  Most  Profitable  Average  Degree  Of  Vacuum  In 
Condenser    Service    is    approximately    as    follows:  (1)    For 
reciprocating  engines,  about  88  per  cent,  of  the  barometer  reading. 
This  corresponds  to  about  26.5  in.  of  mercury  column.     (2) 
For  'turbines  about  95  per  cent,  of  the  barometer  reading.     This 
corresponds  to  about  28.5  in.  of  mercury  column. 

NOTE. — IN  ORDINARY  RECIPROCATING-ENGINE  PRACTICE  it  is  usually 
undesirable  to  carry  a  higher  vacuum  than  about  26.5  in.  of  mercury. 
There  may,  as  hereinafter  specified,  be  several  reasons  for  this:  (1)  If 
the  water  discharged  by  the  condenser  is  to  be  used  for  boiler-feed,  its 
temperature  should  in  many  cases,  for  economic  reasons,  be  higher  than 
that  which  is  due  to  condensation  of  steam  in  a  26.5  in.  vacuum.  The 


286 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


temperature  due  to  this  degree  of  vacuum  will  be  somewhat  less  than 
120  deg.  fahr.  The  cost  of  restoring  heat  to  the  120  deg.  feed-water, 
to  raise  it  to  a  temperature  of  about  210  deg.  fahr.,  which  is  suitable  for 
boiler  feed,  may  make  the  carrying  of  a  high  vacuum  unprofitable.  (2) 
But  even  where  the  supply  of  boiler-feed  water,  and  the  heating  thereof, 
is  independent  of  the  condenser,  a  higher  vacuum  than  about  26.5  in.  is 
rarely  justified  for  reciprocating  engines,  insamuch  as  the  first  cost  of 
the  installation  would  thereby  be  greatly  increased.  The  higher  the 
vacuum,  the  greater  the  cost  (annual  charge)  of  obtaining  each  inch 
increase  of  vacuum.  Considerable  extra  expense  would  be  incurred  by 


E 

3  170 

f 

f 

314J 

j 

c 

s 

<c 

-C 

5: 

§ 

•irurt 

/ 

/ 

f 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

it.  Of  Steam  Consump 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

/ 

3 

1  * 

/ 

/ 

D                19               I 
Vote  u  u  m  In  Inch 

9                2.7                 Z 
5S-Referrec{    To 

30-In.  Barometer 

Absolute    Back    Pressures    In    Lbs.  Per   Sq.In. 

Fio.  289. — Graph  Showing  Relation  Between  Steam  Consumption  And  Condenser 
Vacuum  In  Average  Turbine  Operation.  (The  abscissae  represent  percentages  of 
steam  consumption  with  28-in.  vacuum.)  (Harrison  Safety  Boiler  Works  Catalog.) 

the  additional  precautions  that  would  be  necessary  to  prevent  leakage 
of  air  through  valves,  stuffing-boxes  and  jointed  connections.  (3)  Also, 
the  initial  condensation  in  the  engine  cylinder  (See  the  Author's  STEAM 
ENGINES),  due  to  the  low  temperature  of  the  exhaust  steam,  would  be- 
come so  excessive  as  to  more  than  nullify  the  extra  advantage  gained  in 
lowering  the  back-pressure. 

IN  TURBINE  AND  UNIFLOW-ENGINE  PRACTICE,  however,  the  best 
results  are,  aside  from  considerations  regarding  the  feed-water,  as  noted 
above,  obtained  with  the  highest  vacuum  which  it  is  possible  to  maintain. 
Initial  condensation  (as  is  explained  in  the  author's  STEAM  TURBINES) 
plays  no  part  in  this  case.  Also,  with  turbines,  leakage  into  the  condenser 
can  be  avoided  with  less  difficulty  than  in  reciprocating-engine  practice. 
The  effect  of  variation  in  vacuum  on  turbine  economy  is  shown  graphic- 
ally in  Fig.  289. 


SEC.  329] 


STEAM  CONDENSERS 


287 


329.  Table  Showing  Comparative  Economy  Of  Condensing 
And  Non-Condensing  Operation.  (Compiled  by  the  Inter- 
national Text  Book  Company). 


Steam  consumption  per  indicated 

h.p.  hr.  in  Ib. 

Per  cent. 

Type  of  engine 

Non-condensing 

Condensing 

gained 
by  con- 

t 

densing 

Probable 

Probable 

Probable 

Probable 

limits 

average 

limits 

average 

Simple  high  speed.  . 

40-26 

33 

25-19 

22 

33 

Simple  low  speed.  . 

32-24 

29 

24-18 

20 

31 

Compound   high 

speed  

30-22 

26 

24-16 

20 

23 

Compound   low 

speed  

25-18 

25 

20-13 

18 

25 

Triple  high  speed.  . 

24-17 

20 

18-13 

15 

20 

Triple  low  speed  .  .  . 

27-21 

24 

23-14 

17 

29 

330.  Table  Showing  Steam  Consumption  Of  Condensing 
And  Non-Condensing  Engines.  (From  Gebhardt's  STEAM 
POWER  PLANT  ENGINEERING.) 


Type  of  engine 

Pounds  of  steam 
per  i.h.p.  hr., 
non-condensing 

Pounds  of 
steam  per 
i.h.p.  hr., 
condensing 

Per  cent, 
saving  due  to 
condensing 
operation 

Single  valve  —  simple.  . 
Four  valve  —  simple  .  .  . 
Compound  engine.  .  .  . 

Average—  27  .  63 
Average  —  24  .  06 
Average  —  20.30 

25.7 
19.84 
12.14 

7.0 
17.5 
40.5 

NOTE. — The  performances  shown  in  Table  330  are  for  engines  of  a 
higher  grade  of  construction  than  are  those  shown  in  Table  329.  It  is 
noted  in  Table  330  that  the  per  cent,  of  saving  is  quite  pronounced  with 
the  compound  engine.  This  is  due  to  the  fact  that  the  compound 
engines  are  better  adapted  than  simple  engines  to  handling  the  wide 
temperature  and  pressure  ranges,  which  occur  in  condensing  operation, 
without  excessive  cylinder  condensation  (Sec.  338)  and  other  thermal 
losses  within  the  engine  itself.  For  a  more  complete  discussion  of  the 


288 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


relative    merits    of    condensing    and    non-condensing    operation    under 
different  conditions  see  the  author's  STEAM  ENGINES. 

331.  With  Surface-Condenser  Operation  The  Same   Feed 
Water  May  Be  Used  Repeatedly  (See  Sec.  368).     This  is  an 
important  consideration  where  water  must  be  purchased,  or 
must  be  treated  chemically  to  render  it  suitable  for  boiler  use. 
The  condensate  from  surface   condensers  may,  when  means 
are  employed  to  remove  the  cylinder  oil  therefrom,  be  used 
over  and  over  again  as  boiler  feed.     This  will  tend  to  prevent 
formation  of  scale  in  the  boilers.     Saving  of  fuel  and  reduction 
of  boiler  room  costs  will  thereby  result. 

332.  The   Advantages  And  Disadvantages   Of  Condensing 
Operation  may  be  summarized  as  follows: 


Condensing 


Non-condensing 


Advantages 


Disadvantages 


Decreases  engine  steam  consumption 
20  to  50  per  cent,  in  large  plants. 

Recovers  most  of  the  feed-water  if 
surface  condenser  is  used. 

Feed-water  is  available  at  100  to  120 
deg.  fahr.  unless  very  high  vac- 
uum is  used. 

Decreases  size  of  boiler  installation. 

Increases  power  output  of  a  given 
engine  25  to  95  per  cent. 


Wastes  most  of  the  exhaust  steam 

unless  it  can  be  used  for  heating. 
Must  use  fresh  feed-water  which 

may    be    expensive    to    purify. 
Requires    more     water-purifying 

equipment. 
Feed-water     usually     cold,     and 

must  be  heated  more. 
Requires  larger  boiler  installation. 


Disadvantages 


Advantages 


Requires  additional  equipment,  i.e., 
hot-well,  condenser,  cooling  tower, 
pond  or  source  of  cooling  water, 
vacuum  pump,  circulating  pump, 
condensate  pump,  primary  heater, 
etc. 

Operation  is  more  difficult — requires 
more  intelligent  operators. 

No  steam  available  for  heating. 

Difficulty  of  keeping  joints  tight. 

More  equipment  to  be  kept  in  repair. 


Relatively  low  first  cost. 


Operation  simple — can  be  handled 
by  less  skillful  operators. 

Large  surplus  of  steam  available 
for  heating. 

Small  steam  leaks  do  relatively 
little  harm. 


SEC.  333] 


STEAM  CONDENSERS 


289 


333.  Condensers  May  Be  Classified  Into  Two  General 
Groups:  (1)  Jet  condensers  (Figs.  290  and  291),  in  which  con- 
densation is  by  direct  contact.  That  is,  the  exhaust  steam  and 
the  cooling  water  are  mixed  together.  (2)  Surface  condensers 
(Figs.  292  and  293),  in  which  the  steam  and  cooling  medium 
as  water  or  air,  are  separated  by  metal  walls  or  tubes. 
Heat  is  abstracted  (Sec.  348)  from  the  steam,  through  the 
metal,  by  the  cooling  medium.  The  jet  condensers  will  be 
treated  first  and  then  the  surface  condensers. 


-Water  Pump 


Fia.   290. — Diagram  Of  Elementary  Jet  Condenser. 


334.  There  Are  Three  General  Classes  Of  Jet  Condensers : 

(1)  Standard  low-level  jet  condensers  (Fig.  291),  in  which  the 
water,  steam  and  air  are  exhausted  by  pumps.  (2)  Siphon  jet 
condensers  (Fig.  294),  in  which  the  water,  steam,  and  some- 
times air  are  exhausted  by  a  barometric  column.  (3)  Ejector 
jet  condensers  (Fig.  295),  in  which  the  steam  and  air  are  ex- 
hausted by  the  velocity  or  ejector  effect  of  the  cooling  water. 

NOTE. — Jet  condensers  may  be  further  classified  on  the  basis  of  their 
operation  as  follows:  (1)  Parallel  current  condensers  (Fig.  291)  in  which 
the  condensed  steam,  cooling  water  and  air  flow  in  the  same  direction 
and  collect  at  the  bottom  of  the  condensing  chamber,  whence  they 
are  evacuated  by  a  pump,  barometric  tail  pipe  or  other  means.  These 
condensers  are  used  with  low-vacuum  installations  only.  (2)  Counter- 
current  condensers  (Fig.  296),  in  which  the  condensate  and  cooling  water 
are  taken  off  at  the  bottom  while  the  air  is  removed  at  the  top. 
19 


290 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


335.  The  Cooling  Methods  Employed  With  Surface  Con- 
densers may  be  classified  as  follows:  (1)  Water-cooling,  in 
which  heat  is  abstracted  from  the  steam  by  circulating  water- 
currents.  (2)  Air-cooling,  in  which  the  heat  is  abstracted  and 
carried  away  by  air-currents.  This  method  is  used  only  where 


Pump  For  Discharging  Air- — '' 
Conc/enscrfe  And  tooling-defter 


FIG.  291. — Worthington  Independent  Jet  Condenser. 


water-cooling  is  impracticable,  due  to  an  insufficient  supply. 
(3)  Evaporation  cooling,  in  which  a  cooling  effect  is  produced 
by  the  evaporation  of  streams  of  water  trickling  on  the  outer 
surfaces  of  metal  tubes  through  which  the  exhaust  steam  is 
made  to  flow. 


SEC.  335] 


STEAM  CONDENSERS 


291 


Water  Outlet 


'Steam 


Condensate 
Pump 


From 

Dripping 

Into 

To  Air   Air  Pipe 

Pump 


FIG.  292. — Elementary  Double-Flow  (Two-Pass)  Surface  Condenser. 


Condenser  She!l\ 


.•Suction 
•    Inlet 


Exhaust- 


Circulating  Water 

Leaves  Condenser-... 

:i^i 


.--Delivery  Valves- 

f,.--Sucf/bn  Valves- •  • .. . .  J 


''•Wet-Vacuum  Pump       Steam-Power  Cylinder-'    Circulating  Pump--' 


FIG.  293. — Typical  Surface  Condenser  With  Combined  Circulating  And  Wet-Vacuui 
Pumps  Of  Reciprocating  Type. 


292 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


NOTE. — Water-cooling  is  used  almost  exclusively  in  the  operation  of 
steam  condensers.  All  subsequent  mention  of  condensers  in  this  book 
will,  therefore,  be  limited  to  the  water-cooled  type. 


-Cooling-Water  Inlet 


.-Condensing 
Water 


'""••Relief  Valve 


Drain- 


Hot 

Well > 


•Overflow 


FIG.  294.— Diagrammatic  Sectional  View  Of  Buckley  Siphon  Condenser  And 
Connections. 


33 5 A.  The  Circulation  Of  The  Water  In  Surface  Condensers 

may  be :  (1)  A  single  flow,  as  where  the  water  passes  (Fig.  297) 
only  once  through  the  tubing.  That  is,  the  water  flows  into 
the  condenser  at  one  end,  through  the  tubes,  and  out  at  the 
other  end.  (2)  A  double  flow,  as  where  a  division-plate  (Fig. 


SEC.  335] 


STEAM  CONDENSERS 


293 


293)  is  inserted  so  that  the  water  passes  twice  through  the 
tubing.     That  is,  the  water  passes  first  through  one  portion 


ExhotusfSfeorm-.^ 
Entrance 


Cooling 

Water 

.•Entrance 


''-Aspirotti'ny 
Cones 


.-Tall  Pipe 
\ 


FIG.   295. — Koerting  Multi-Jet  Ejector  Condenser  (For  Large  Capacities). 

of  the  tubes  and  returns  through  the  remaining  tubes,  thereby 
the  water  travels  twice  the  length  of  the  tubes.  (3)  A  multi- 
flow  (Fig.  298),  as  where  two  or  more  division  plates  are  in- 


294 


STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 


Vacuum  ,--4/r  Pump 

/Breaker  S'  Connection 


Fia.  296.— Wheeler  Rectangular  Rain  Type  Of  Low-Level  Jet  Condenser  For  High 
Vacuum  Service — Especially  Adapted  For  Use  With  Turbines. 


.•Exhaust  Steam  A  Discharge  Outlef\ 

'" 


Drain-''        Cono/errser---' 


FIG.  29  7. — Baragwanath 
Single  Flow  Surface  Con- 
denser. 


...-•Hanolhotes      Exhaust  Inlet-  ...... 


Outlet*. 


"~* -Diaphragms 

iHet  k  AirArtCondensatePump 

Fia.     298.— A    Multiflow    Surface    Condenser. 
(This  is  a  "four-pass"  condenser.) 


SEC.  3361  STEAM  CONDENSERS  295 

serted  so  that  the  water  makes  three  or  more  passes  through 
the  tubing. 

NOTE. — Most  surface  condensers  are  of  the  double-flow  or  multiflow 
type.  Single-flow  condensers  are  rarely  used.  Double-flow  or  multi- 
flow  condensers  are  used  where  high  vacua  are  required,  as  in  turbine 
operation. 

336.  The     Operation    Of    A     Standard    Low-Level    Jet 
Condenser   (Fig.   291)   is  as  follows:  A  direct-acting  steam 
pump,  to  which  the  condenser  is  attached,  removes  the  con- 
densate,  cooling  water  and  entrained  air  through  a  common 
cylinder,  P.    The  cooling  water  enters  at  W.    The  quantity  of 
water  is  controlled,  in  accordance  with  the  quantity  of  ex- 
haust steam  to  be  condensed,  by  the  valve  S,  which  is  ad- 
justed by  means  of  the  hand-wheel  H.     The  valve  S  is  so 
shaped  as  to  deliver  the  water  in  the  form  of  a  spray.     The 
exhaust  steam,  which  enters  at  E,  mixes  with  the  spray  of 
cooling  water  in  the  chamber  B.     The  mingled  current  of 
condensate  and  cooling  water  is  then  discharged  by  the  pump. 

NOTE. — Assuming  that  the  cooling-water  supply  is  adequate,  the 
vacuum  will  be  maintained  in  a  jet  condenser  as  long  as  the  pump  keeps 
the  condensing  chamber  free  from  water  and  air.  The  degree  of  vacuum 
in  a  jet  condenser  depends  upon  the  water-temperature  and  the  quantity 
of  entrained  air.  The  cooling  water  is,  generally,  drawn  in  by  the  vacuum, 
instead  of  being  delivered  by  gravity.  This  serves  to  safeguard  the 
engine  in  case  of  accidental  stoppage  of  the  pump.  Should  the  pump 
suddenly  stop,  enough  water  would  almost  instantly  accumulate  to 
submerge  the  spray  valve  S.  However,  since  insufficient  water  surface 
to  condense  the  steam  would  then  be  presented,  the  vacuum  would 
immediately  be  broken.  Hence,  flooding  of  the  condenser  and  wrecking 
of  the  engine  thereby  could  not  occur.  The  engine  would  then  exhaust 
to  the  atmosphere  (Fig.  299)  through  the  relief  valve. 

337.  To  Put  A  Standard  Jet  Condenser  In  Service,   pre- 
paratory to  starting  the  engine  to  which  it  is  attached,  the 
injection  valve  V  (Fig.  299)   should  first  be  opened.     The 
pump  should  then  be  brought  up  to  its  regular  running  speed. 
The  engine  may  then  be  started.     If  the  pump  suction  is  not 
sufficient  to  raise  the  water  from  the  well  the  condenser  must 
be  primed  with  a  small  amount  of  cold  water  introduced 
through  valve  E.     Adjustment  of  the  spray  valve  S  may  then 
be  necessary  to  produce  the  required  degree  of  vacuum. 


296 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


NOTE. — If  two  jet  condensers  are  to  be  used  on  the  same  exhaust  line, 
they  should  be  connected  independently  (Fig.  300,  pipes  D).     If  their 


Steam  Supply  Pipe--*- 


Engine 
C/fodv- 


FIG.  299. — Installation  Of  Jet  Condenser  With  Reciprocating  Engine. 

circulating  systems  are  connected  in  series  as  by  pipe  A,  the  arrangement 
will  be  unsatisfactory.  Condenser  1  will  use  water  at  a  lower  tempera- 
ture than  condenser  2,  and  the  vacuum  will  therefore  be  higher  in  1. 


ufxhcryt  Main       "<"  ^-'-Proper  Cbh'nec'f/oh' 
•P/pe  Line  Below  Floor  Of  Suction  Pipe 

Fia.  300.— Showing  Jet  Condensers  Connected  In  Parallel  And  With  Injection-Water 
Connections  In  Series. 


If  the  vacuum  is  equalized  in  the  two  condensers  by  partially  closing 
the  exhaust  steam  valve  to  condenser  2,  then  nearly  all  the  condensation 
will  take  place  in  1. 


SEC.  338]  STEAM  CONDENSERS  297 

338.  To  Stop  A  Standard  Jet  Condenser,  after  closing  the 
throttle  of  the  engine  to  which  it  is  attached,  the  injection 
valve    V    (Fig.    299)    should   first   be   closed.     The   vacuum 
breaker,  D,  should  then  be  opened.     The  pump  may  then  be 
stopped. 

NOTE. — The  momentum  of  the  flywheel  will  cause  an  engine  to  con- 
tinue in  motion  for  several  seconds  after  the  throttle  is  closed.  During 
this  interval  the  movement  of  the  piston  will  tend  to  produce  a  pumping 
effect.  Hence,  a  slug  of  water  may  be  drawn  into  the  engine  cylinder  if 
th,e  condenser  pump  is  shut  down  before  the  injection  water  is  shut  off 
and  the  vacuum  broken.  This  may  occur  where  the  engine  cylinder  is 
less  than  about  22  ft. — which  is  the  practical  maximum  suction  lift  in 
pump  operation  (See  Sec.  1)  — above  the  level  of  the  condenser,  pump,  or 
other  source  from  which  the  water  might  be  sucked  into  the  engine 
cylinder. 

339.  The  Operation  Of  A  Siphon  Or  Barometric  Jet  Con- 
denser (Fig.  294)  is  as  follows:  The  cooling  water  which  is 
supplied  by  the  pump  enters  at  E  and  passes  downward  around 
the  exhaust  nozzle,  N,  in  a  thin  conical  film.     The  exhaust 
steam  from  N  is  condensed  within  this  hollow  cone  of  falling 
water,    thus    creating    the    desired    partial    vacuum.     The 
condensate  and  cooling  water  are  discharged  from  the  condens- 
ing chamber  C  by  a  barometric  tail-pipe  T.     The  lower  end 
of  the  tail-pipe  is  submerged  in  a  hot- well,  H.     In  flowing 
through  the  neck,   or  constricted  passage,  K,  the  mingled 
current  of  condensate  and  cooling  water  acquires  sufficient 
velocity  to  draw  out  the  air  which  may  be  entrained  with  the 
steam. 

NOTE. — The  chief  purpose  of  this  condenser  arrangement  is  to  obviate 
liability  of  damage  to  the  engine  by  water  being  drawn  from  the  tail-pipe 
into  the  condensing  chamber  and  thence  to  the  exhaust  pipe.  The  level 
of  the  water  in  the  hot-well,  H}  is  at  least  35  ft.  below  the  condensing 
chamber. 

Atmospheric  pressure  cannot  sustain  a  column  of  water  having  a  height 
exceeding  34  ft.  Hence  it  is  impossible  for  water  to  get  above  the 
nozzle  N.  If  the  level  of  the  injection-water  supply  is  not  more  than 
about  20  ft.  below  the  inlet,  E,  to  the  condenser,  the  siphoning  action  of 
the  tail-pipe  will  suffice  to  raise  the  water.  The  pump  may  then  be 
dispensed  with  after  the  vacuum  has  been  formed.  But  if  a  lift  of  20  ft. 
is  to  be  exceeded  the  pump  must  be  run  continuously. 


298 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


EXPLANATION. 
priming  valve  S. 


Injection  Wetter 


340.  Siphon  Jet  Condensers  May  Be  Started  And  Operated 
Without  A  Pump  (Fig.  301)  if  the  injection  water  is  to  be 
lifted  less  than  20  ft.  In  such  cases,  the  full  siphoning  action 
of  the  tail-pipe,  due  to  gravity,  is  produced  by  a  double-stage 
operation. 

Water  is  first  admitted  to  the  tail-pipe  through  the 
In  falling  through  the  tail-pipe,  this  current  of  water 
draws  out  the  air.  It  thus  pro- 
duces a  sufficient  vacuum  in  the 
condenser  and  upper  part  of 
the  tail-pipe  to  draw  the  cooling 
water  to  the  condenser  through 
the  injection  pipe  W.  The 
priming  valve,  S,  is  then  closed. 
The  flow  of  injection  water  is 
regulated  by  means  of  the 
valve  R. 

NOTE.  —  Sometimes,  if  the  in- 
jection supply  is  to  be  lifted  to 
a  height  of  about  20  ft.,  a 
priming  pipe,  and  valve  P,  lead- 
ing from  the  boiler-feed  pump, 
must  be  installed.  The  boiler- 
feed  pump  may  then  be  uti- 
lized as  an  aid  in  starting  the 
condenser. 

341.  The  Operation  Of 
An  Ejector  Jet  Condenser 

(Fig.  295)  is  as  follows: 
The  cooling  water,  entering 
at  W,  is  supplied  under  a 
sufficient  head,  due  either  to  gravity  or  pump  action,  to  send 
it  through  the  constricted  neck,  X,  and  into  the  tail-pipe, 
T,  at  very  high  velocity.  The  exhaust  steam  enters  at  S 
and  is  condensed  by  contact  with  the  jets  of  cooling  water. 
The  aspiratory  or  suction  effect,  due  to  the  high  velocity  of  the 
jets,  draws  the  entrained  air  through  the  openings  between  the 
aspirating  cones,  A,  whence  it  is  picked  up  and  ejected  with 
the  mingled  current  of  condensate  and  cooling  water. 

342.  The  Requisite  Size  Of  Jet  Condensers  is  generally 
determined  in  accordance  with  the  records  of  actual  experience 
in  condenser  operation.  Numerous  empirical  formulae  based 


FIQ.  301. — Apparatus  For  Starting  A  Siphon 
Condenser  When   No  Pump  Is  Used. 


SEC.  343]  STEAM  CONDENSERS  299 

on  practice  have  been  developed.  A  satisfactory  formula, 
from  MARK'S  HANDBOOK,  is  as  follows: 

(88)  V  =  0.00143Wa  +  8.25  (cubic  feet) 

Wherein  V  —  the  volume  of  the  condenser,'in  cubic  feet.  W8 
=  the  weight  of  steam,  in  pounds,  to  be  condensed  per  hour. 

EXAMPLE. — A  standard  jet  condenser  is  required  for  a  2,000-h.p. 
engine,  the  steam  consumption  of  which  will  be  approximately  17  Ib. 
per  h.p.  per  hr.  What  should  be  the  volume  of  the  condenser? 

SOLUTION.— By  For.  (88),  V  =  0.00143WS  +  8.25  =  (0.00143  X  17  X 
2,000)  +  8.25  =  56.9  cu.  ft. 

NOTE. — THE  VELOCITY  OF  THE  EXHA.UST  STEAM  ENTERING  A  JET 
CONDENSER  should  be  approximately  600  ft.  per  sec.  The  exhaust  steam 
inlet  should  be  proportioned  to  produce  this  velocity. 

NOTE. — THE  VELOCITY  OF  THE  CONDENSATE  AND  COOLING  WATER 
ISSUING  FROM  A  JET  CONDENSER  should  be  approximately  as  follows: 
(1)  For  a  standard  jet  condenser  (Fig.  291)  5  ft.  per  sec.  (2)  For  a  siphon 
condenser  (Fig.  294)  5  to  10  ft.  per  sec.  (3)  For  an  ejector  condenser 
(Fig.  295)  15  to  20  ft.  per  sec.  The  outlet  from  the  bell  or  condensing 
chamber  should  be  so  proportioned  as  to  produce  these  velocities. 

343.  The  Quantity  Of  Cooling  Water  Required  For  Jet 
Condensers  depends  upon  the  following  factors:  (1)  The  degree 
of  vacuum  required.     (2)  The  heat  of  the  exhaust  steam.     (3) 
The  effectiveness  with  which  the  steam  and  water  are  mixed.     (4) 
The  quantity  of  air  entrained  with  the  steam.    (5)  The  general 
efficiency  of  the  condensing  equipment.     The  exhaust  from  an 
engine  generally  contains  considerable  moisture.     For  practi- 
cal purposes,  however,  it  is  sufficiently  accurate  to  assume  that 
it  consists  entirely  of  dry,  saturated  steam. 

344.  To  Compute  The  Quantity  Of  Cooling  Water  Required 
For  Jet  Or  Surface  Condensers  the  following  formula  may  be 
used: 

TT  m         I     oo 

(89)  Ww  =  Ws      Tf2_r*  (pounds) 

Wherein:  Wu,  =  the  weight  of  water,  in  pounds  per  hour, 
which  is  required  to  condense  and  cool  the  exhaust  to  a  given 
discharge  temperature.  Ws  =  the  weight  of  steam  to  be  con- 
densed per  hour,  in  pounds.  H  =  the  quantity  of  heat, 
above  32  deg.  fahr.,  in  British  thermal  units,  in  1  Ib.  of  dry, 
saturated  exhaust  steam  at  the  condenser  pressure,  as  given 


300  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

in  Table  346.  T/i  =  the  temperature  of  the  entering  cooling 
water,  in  degrees  Fahrenheit.  T/2  =  the  temperature  of  the 
discharged  cooling  water,  in  degrees  Fahrenheit.  T/c  =  tem- 
perature of  the  condensate,  in  degrees  Fahrenheit.  (This 
temperature  is  the  same  as  T/2  in  jet  condensers.) 

EXAMPLE. — The  vacuum  gage  of  a  jet  condenser  registers  26  in.  of 
mercury.  The  barometer  registers  an  atmospheric  pressure  of  29.4  in. 
of  mercury.  The  cooling  water  enters  the  condenser  at  a  temperature 
of  70  deg.  fahr.  The  temperature  of  the  discharge  is  105  deg.  fahr.  The 
steam  consumption  of  the  engine  is  30,000  Ib.  per  hr.  What  quantity  of 
cooling  water  is  required? 

SOLUTION. — The  absolute  condenser  pressure  =  29.4  —  26  =  3.4  in.  of 
mercury.  Hence,  if  the  barometer  reading  (Table  346)  were  30  in.  of 
mercury,  the  vacuum  gage  would  show  30  —  3.4  =  26.6  in.  of  mercury. 
By  Table  346,  the  total  heat  in  the  steam  above  32  deg.  fahr.  correspond- 
ing to  a  vacuum  of  26.6  in.  of  mercury  =  1,112.2  B.t.u.  per  Ib.  Hence, 
by  For.  (89),  W«  =  WS(H  -  T/c  +  32)/(T/2  -  Tfl}  =  30,000  X  (1,112.2 
-  105  +  32)  -T-  (105  -  70)  =  890,700  Ib.  per  hr. 

NOTE. — THE  TEMPERATURE  OF  THE  WATER  DISCHARGED  FROM  A 
JET  CONDENSER  is  always  lower  than  the  temperature  (Table  346) 
which  is  due  to  the  condenser  pressure.  In  high  class  installations  it 
may  be  only  5  deg.  fahr.  below  this  temperature.  With  poorly  designed 
condensers  it  may  be  20  deg.  below.  '  But  the  average  difference  is  from 
10  to  15  deg. 

345.  The  Operation  Of  A  Surface  Condenser  is  as  follows: 
The  cooling  water  is  pumped  through  the  tubes  (Fig.  293)  by 
the  circulating  pump  P.  The  exhaust  steam  enters  at  S.  The 
condensate  and  air  are  drawn  out  by  the  vacuum  pump  C. 
The  cooling  water,  if  admitted  at  the  bottom,  will  first  act  upon 
that  portion  of  the  steam  which  is  at  the  lowest  temperature. 
This  is  conducive  to  effective  transfer  of  heat  from  the  steam 
to  the  water.  It  is  called  the  counterflow  principle. 

NOTE. — HEAT  TRANSFERENCE  IN  SURFACE  CONDENSERS  MAY  BE 
IMPROVED  by  preventing  the  condensate  which  forms  on  the  upper  tubes 
from  falling  on  the  lower  tubes.  Baffles,  or  rain-plates,  are  sometimes 
employed  for  this  purpose.  Condensers  so  equipped  are  called  dry-tube 
condensers.  By  keeping  the  lower  tubes  comparatively  dry,  condensa- 
tion of  the  steam  in  the  lower  half  of  the  condenser  proceeds  more  rapidly 
than  it  otherwise  would.  Films  of  water  enveloping  the  tubes  serve  to 
insulate  them. 


SEC.  346] 


STEAM  CONDENSERS 


301 


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302 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


347.  The  Tubes  And  Tube-Sheets  Of  Surface  Condensers 
are,  generally,  made  of  such  metals  as  are  best  adapted  to 
resist  the  corrosive  action  of  the  waters  which  are  available 
for  cooling  purposes.  Where  fresh  water  is  used  the  tubes 
may  be  of  brass,  bronze,  copper,  aluminum-bronze,  or  Muntz 
metal.  Where  salt-water  is  used,  tubes  made  of  Admiralty 
metal  are  preferred.  This  is  an  alloy  containing  70  per  cent, 
of  copper,  29  per  cent,  of  zinc  and  1  per  cent  of  tin.  The  tube- 
sheets  are  generally  made  of  brass  or  Muntz  metal.  The  shell 
and  fittings  are  commonly  made  of  cast  iron. 

NOTE. — The  sizes  of  condenser  tubes  in  common  use  are  £s-in.,  M-in., 
and  1-in.  outside  diameter.  The  corresponding  thicknesses  are  20,  18 
and  16  Birmingham  wire  gage.  Fig.  301 A  shows  one  way  in  which  tubes 
are  fastened  in  the  tube  sheets  or  tube  head. 


Glancf* 


FIG.  301A. — Worthington  Standard  Condenser-Tube  Gland. 


348.  The  Quantity  Of  Heat  Which  The  Cooling  Water 
Will  Abstract  From  Steam  In  Surf  ace -Condenser  Operation 
may  be  computed  by  the  following  formula : 

(90)  Ht  =  WS(H  -  Tfc  +  32)  (British  thermal  units  per  hour) 

Wherein  Ht  =  the  quantity  of  heat  given  up  by  the  steam, 
in  British  thermal  units  per  hour.  W8  =  the  weight  of  steam 
condensed,  in  pounds  per  hour.  H  =  the  total  heat,  above 
32  deg.  fahr.,  in  British  thermal  units,  in'l  Ib.  of  the  exhaust 
steam  at  the  condenser  pressure,  as  given  in  Table  346.  T  fc  = 
the  temperature,  in  degrees  Fahrenheit,  of  the  condensate 
leaving  the  condenser. 

EXAMPLE. — The  steam  consumption  of  a  1,000-h.p.  engine,  exhausting 
into  a  surface-condenser,  is  18  Ib.  per  h.p.  hr.  The  average  vacuum-gage 
reading  is  25.4  in.  of  mercury.  The  average  atmospheric  pressure,  as 


SEC.  349]  STEAM  CONDENSERS  303 

shown  by  the  barometer,  is  30  in.  of  mercury.  The  temperature  of  the 
discharged  condensate  is  120  deg.  fahr.  How  much  heat  is  given  up  by 
the  cooling-water? 

SOLUTION. — By  Table  346,  the  total  heat,  above  32  deg.  fahr.,  in  the 
steam,  for  a  25.4-in.  vacuum  with  a  30-in.  barometer,  is  1,116.9  B.t.u. 
per  Ib.  By  For.  90,  Ht  =  WS(H  -  Tfc  +  32)  =1,000  X  18  X  (1116.9  - 
120  +  32)  =  18,520,000  B.t.u.  per  hr. 


349.  The  Water-Cooling,  Or  Tube  Surface,  Required  In  A 
Surface  Condenser  may  be  computed  by  the  following 
formula : 

(91)  At  =  — H*     ,    ,  x  (square  feet) 


U 


1T 

Vfs 


Wherein  A/  =  the  water-cooling,  or  tube  surface,  in  square 
feet.  Ht  =  the  quantity  of  heat  to  be  given  up  by  the  steam, 
in  British  thermal  units  per  hour,  as  computed  by  For.  (90). 
T/s  =  the  temperature  of  the  steam,  in  degrees  Fahrenheit, 
as  given  in  Table  346.  U  =  a  constant  from  Table  350  = 
B.t.u.  transferred  per  square  foot  per  hour  per  degree  tempera- 
ture difference  between  the  water  and  the  steam.  T/i  and 
T/2  =  ,  respectively,  the  initial  and  final  temperatures  of  the 
cooling  water  in  degrees  Fahrenheit. 


EXAMPLE.- — The  heat  to  be  abstracted  from  the  exhaust  steam  entering 
an  ordinary  type  of  standard  surface  condenser,  as  computed  by  For. 
90,  amounts  to  18,000,000  B.t.u.  per  hr.  The  average  vacuum-gage 
reading  is  assumed  to  be  25.1  in.  of  mercury.  The  average  atmospheric 
pressure,  as  shown  by  the  barometer,  is  assumed  to  be  30  in.  of  mercury. 
The  cooling-water  is  assumed  to  enter  at  a  temperature  of  55  deg.  fahr. 
and  emerge  at  a  temperature  of  100  deg.  fahr.  How  much  tube-surface 
is  required? 

SOLUTION. — By  Table  346,  the  temperature  of  the  steam,  for  a 
25.1  in.  vacuum  with  a  30-in.  barometer,  is  133  deg.  fahr.  By  Table  350, 
the  coefficient,  C7,  of  heat  transference  is  250.  By  For.  (91),  Af  = 
Ht/(U[Tf.~  M(?Vi  +  Tft)]}  =  18,000,000  -f-  {250  X  [133  - 
M  (55  +  100)])=  1,297  sq.ft. 


304 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


350.  Table  Of  Coefficients  Of  Heat  Transference  (U,  For. 
91)  In  Surf  ace -Condenser  Operation. 


Type  of  surface 
condenser 

Velocity  of  cooling 
water,  in  feet 
per  second 

Value  of  U,  in  B.t.u. 
per  sq.  ft.  per  deg. 
temp.  dif.  between 
water  and  steam 

Ordinary,  old  style, 
standard  type 

1  to  2 

250 

Modern,  dry-tube  type 

4  to  5 

600 

351.  The  Value  Of  The  Heat  Transference  Coefficient,  U 
(Table  350),  may  range  from  1,000  to  3  between  different 
areas  of  the  tube-surface  in  the  same  condenser.     The  values 
given  in  Table  350  are  average  values.     From  tests  made  by 
Prof.  Josse,  of  the  Royal  Technical  School  at  Charlottenburg, 
it  was  found  that  the  value  of  U  is  affected  principally  by  the 
following    factors:  (1)  The    material,    thickness,    shape    and 
cleanliness   of  the   tubes.     (2)   The   water-velocity   through   the 
tubes.     (3)  The  steam-velocity  over  the  tubes.     (4)  The  quan- 
tity of  condensate  adhering  to  the  tubes. 

NOTE. — The  results  of  actual  practice  have  demonstrated  that  sur- 
face condensers  of  the  ordinary  standard  type,  when  attached  to  engines 
using  20  Ib.  of  steam  per  h.p.hr.,  and  operating  with  a  26-in.vacuum,  re- 
quire about  2  sq.  ft.  of  tube-surface  per  engine  horsepower.  Also,  that 
dry-tube  multiflow  condensers,  when  attached  to  turbines  using  15  Ib. 
of  steam  per  k.w.hr.,  and  operating  with  a  28.5-in.  vacuum,  require 
about  2  sq.  ft.  of  tube-surface  per  kilowatt  developed  by  the  turbine. 
Condenser  practice  in  general  indicates  that  from  1.25  to  2.5  sq.  ft.  of 
tube-surface  per  kilowatt  are  required  for  large  modern-type  installa- 
tions, while  from  2  to  4  sq.  ft.  per  kilowatt  are  required  for  the  smaller 
installations  of  ordinary  standard  equipment. 

352.  The   Temperature   "Drop"   In    Surface   Condensers 
means  the  difference  in  temperature  between  the  entering 
steam   and   the    discharged   cooling   water.     With   ordinary 
standard    surface    condensers  of  the   single-flow  or  double- 
flow  type,  the  temperature  "drop"  ranges  usually  from  10 
to   20   deg.    fahr.     With   high-vacuum   multi-flow    dry-tube 


SEC.  353] 


STEAM  CONDENSERS 


305 


condensers,  temperature  drops  of  1  to  5  deg.  fahr.  have  been 
obtained.  The  temperature  difference  between  the  conden- 
sate  and  discharged  cooling  water  is  usually  5  to  10  deg.  fahr. 
353.  The  Classes  Of  Pumps  Used  In  Connection  With  A 
Condenser  are:  (1)  Circulating  pumps,  or  pumps  used  for 
forcing  water  through  the  tubes  of  surface  condensers;  or 
furnishing  water  to  barometric  or  ejector-jet  condensers;  or 
removing  water  from  jet  condensers  having  dry-air  pumps. 
(2)  Wet  vacuum  pumps,  or  pumps  used  for  pumping  both 
condensate  and  air  from  jet  or  surface  condensers.  Wet  air 
pumps  for  jet  condensers  handle  the  injection  water  also  and 


FIG.   302. — Typical   Installation  Of  Turbine  With  High- Vacuum  Jet  Condenser  And 
Pumps  With  10,000  Kw.  Unit. 

are  sometimes  called  simply  condenser  pumps.  (3)  Condensate 
pumps,  or  pumps  used  with  surface  condensers  to  pump  the 
condensed  steam  only,  to  a  heater  or  receiver — usually  for  use 
as  boiler  feed.  (4)  Dry  vacuum  or  air  pumps,  or  pumps  used 
for  removing  air  only,  from  jet  or  surface  condensers.  (5) 
Hot-well  pumps,  or  pumps  used  for  pumping  the  hot  water 
from  a  hot  well  usually  to  a  feed-water  heater. 

354.  The  Types  Of  Pumps  Used  As  Condenser  Auxiliaries 
are:  (1)  Direct-acting  steam  pumps  (Fig.  293).  These  are  used 
chiefly  in  reciprocating  engine  plants  as  wet  vacuum  pumps, 
circulating  pumps  or  condensate  pumps.  (2)  Rotative  or 
crank-action  pumps,  steam  or  power  driven  (Fig.  302) .  These 
20 


306  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

.-•Alternator  _,       Ste&m  Turbine-*^ 


P^fT.  ;•.  -  'Pump  :  ^  •  •/ .   : ._• . 


-^Er±^^-L~:~"5<J^f''?n  T°  '  -'•"'-•-'-'•'-"    •'••'••"'  :f  ' .  ' 
^^^•C'fculcti-ing- Pump  :'•'/•  :  .;•'.'-'  •'   ;*  •    .". 


FIG.  303. — Turbine  With  Westinghouse-Leblanc  Surface  Condenser.      (The  equipment 
shown  has  been  superseded  by  more  modern  designs.) 


Water- Pi's  fans- — -*-;"ii 
PocketectAfc 

Air 
Inlet 


/•Potating  Impeller 
"siuas'Of-. 


\ 


Hurling--- 
Water  In  let 


f^-  5tortionctry\ 
_  $  Guide  Vctney 

Compression  Channel-"' 

Fio.  304.— Illustrating  Principle  Of  'Alberger  Hurling-Water — Centrifugal— Air 
Pump.  (As  the  impeller  revolves,  it  throws  streams  of  water  out  between  its  blades. 
Each  time  a  stream  of  water  passes  a  compression  channel,  a  small  amount  or  "slug"' 
of  water  is  thrown  up  the  channel  with  considerable  force.  Air,  which  is  admitted 
between  the  impeller  and  the  channels,  is  caught  between  the  slugs  of  water  and  carried 
out  with  them.) 


SEC.  3541 


STEAM  CONDENSERS 


307 


are  used  chiefly  for  dry-vacuum  pumps  in  either  turbine  or 
reciprocating  engine  plants.  Crank-action  power  pumps  are 
occasionally  used  for  circulating  and  wet-air  pumps,  but  steam 
drives  are  more  common  because  the  exhaust  steam  from 


..—-Impeller . 


.-Diffusion  Chambers 


Water  Inlet--'  '"'-Discharge  Outlets--"' 

FIG.  305. — Alberger  Hurling- Water  Air  Pump. 

the  drives  is  usually  needed  for  feed-water  heating  in  condens- 
ing plants.  (3)  Centrifugal  pumps  (Figs.  302  and  303). 
These  are  the  most  commonly  used  type  of  circulating  and 
condensate  pumps  in  modern  installations  of  medium  and  large 


Main  Exhaust 


Main 
.'Condenser 


.-•Air  Pump 


Water 


FIG.  306. — Parsons  Vacuum  Augmenter. 


capacity.  (4)  Hurling-water  pumps  (Figs.  303,  304  and  305), 
sometimes  called  hydro-centrifugal  pumps.  These  are  used  as 
dry-vacuum  pumps  chiefly  in  turbine  installations  where  the 
vacuum  is  high  and  the  volume  of  air  to  be  handled  is  relatively 


308 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


small.  (5)  Jet  pumps  or  ejectors  (Figs.  306  and  306 A.). 
These  are  used  for  increasing  vacuum  or  are  built  as  two  and 
three-stage  ejectors  for  high- vacuum  pumping  service. 

Strainer  Cage  • 

Nozzles 
.•(First  Stage) 


From 
Boifer. 


Discharge 
—  Opening 


Fia.  306.A. — Wheeler  Two-Stage  "Radojet"  Air  Pump  Without  Inter-condenser. 
(Air  and  water  vapor  are  drawn  from  the  condenser  in  through  the  suction  chamber,  E, 
by  steam  issuing  from  the  nozzles,  D,  at  a  velocity  of  about  3,000  ft.  per  sec.  The 
mixture  is  discharged  into  the  diffuser,  F,  whence  it  is  led  to  the  double  passage,  G. 
When  an  intercondenser  is  employed,  the  mixture  passes  from  F  to  the  interconderiser 
where  the  steam  is  condensed  and  from  which  the  air  is  led  to  G.  Steam,  delivered 
through  nozzle  throat,  H,  strikes  nozzle  point,  J,  and  forms  a  thin  sheet  issuing  outward 
through  K  and  drawing  air  from  G  into  the  volute,  L,  whence  the  steam  and  air  may  be 
discharged  into  the  atmosphere  or  into  a  properly-vented  feed-water  heater.) 

355.  The  Advantages  Of  Centrifugal  Pumps  For  Condenser 
Circulating  Or  Condensate  Pumps  are:  (I)  Low  first  cost.  (2) 
Compactness.  (3)  Absence  of  valves  and  pistons.  (4)  High 


SEC.  355] 


STEAM  CONDENSERS 

Pump 


309 


<  Piston — ^«®^   i     »\m 

\         .     ^*SskMjji&rtnfrance  Of 

•^  Dram- ^SsSr^  Wafer  And  Air  To 

'-Entrance  To  Pump  Cylinder 


Fio.  307.— Sectional  View  Of  Wheeler-Edwards  Combined  Condensate  And  Air  Pump. 


;£>ry  Air  Pump 
\  Connection 


Supplementary  Injection  Wafer 
Connection  For  Cooling  Air  And 


FIG.  308. — Condensing  Chamber  Of  Alberger  Barometric  Condenser,  Showing  Dry  Air 

Pump  Connection. 


310 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


speed.  This  permits  their  being  driven  through  direct  shaft 
connection  with  electric  motors  and  steam  turbines.  In  fact 
(see  Sec.  Ill)  centrifugal  pumps  are  inherently  high  speed 
machines  which  renders  them  especially  adaptable  for  being 
driven  by  motors  or  steam  turbines  which  are  also  inherently 
high-speed  machines.  The.  same  advantages  apply  to  hurling- 
water  pumps  as  compared  to  piston  pumps  for  dry-vacuum 
pumps. 


•Piston  Rod 


Connection  To 

Water         Condenser 
Jacket-.       (Air And  Vapors 
Pass  ToCy Under 
Ports  Through 
Annular  Passage 
Around  Cylinder) 


Rotary 

Admission 

Valve 


Automatic 
Delivery  Valve-- 


.-•Dr'xhcirge 
Outlet 


FIG.  309.— Wheeler  Dry-Vacuum  Pump  (Single- Valve  Type).  (The  function  of  the 
rotary  admission  valve  is:  (1)  To  connect  the  discharge  outlet  and  inlet  port  alternately 
•with  opposite  ends  of  the  pump  cylinder.  (2)  To  release  the  compressed  air  in  the  clearance 
space  at  one  end  of  the  cylinder  through  the  transfer  passage  to  the  other  end  of  the  cylinder. 
Compressed  air  in  the  clearance  space  is  in  this  way  released  into  the  other  end  of  the 
cylinder  instead  of  back  into  the  suction  pipe  where  it  would  tend  to  decrease  condenser 
vacuum.) 


NOTE. — IN  ORDER  To  SECURE  A  HIGH  VACUUM  WITH  PISTON  PUMPS 
it  is  essential  that  the  clearance  volume  of  the  air-pump  (Fig.  307) 
should  be  kept  as  low  as  possible.  Also,  to  avoid  the  use  of  inconven- 
iently large  pumps  the  mixture  to  be  handled  should  be  cooled  to  the 
lowest  practicable  temperature.  Sometimes  the  air  is  re-cooled  by  the 
incoming  condensing  water  (Fig.  308).  In  some  cases  a  steam  jet 
(Fig.  306)  is  used  to  partially  compress  the  air  before  it  goes  to  the 
pump.  In  others  a  special  valve  (Fig.  309)  is  used  for  reducing  the  pres- 
sure in  front  of  the  piston,  at  the  end  of  the  delivery  stroke,  to  the 
condenser  pressure.  This  is  to  obviate  the  loss,  which  would  otherwise 


SEC.  355A] 


STEAM  CONDENSERS 


311 


result  from  expansion  of  a  portion  of  the  compressed  air  down  to  the 
suction  pressure,  when  the  piston  begins  a  stroke. 

Rotatory  pumps  (hurling  or  hydro-centrifugal  pumps)  using  slugs  of 
water  (Fig.  304)  as  pistons  are  sometimes  used  where  very  high  vacua 
are  required. 

355A.  A  Modern  Westinghouse  Turbine -Generator-Sur- 
face-Condenser Installation  is  shown  in  Fig.  309A.  The 
turbine,  T,  is  connected  to  a  Le  Blanc  surface  condenser,  C, 
by  an  expansion  joint,  X,  and  a  short  connecting  piece,  J. 


.  ^Air  Pump  D/schorrae 


nv--; fn/ef  Cao/ingf  Wetter  Tunnel 


FIQ.  309 A. — Westinghouse-LeBlanc    Surface   Condenser    Installed    For   Service   With 
Turbo-Generator  Unit. 

The  expansion  joint  is  necessary  to  properly  protect  the  tur- 
bine and  condenser  shell  from  excessive  stress  due  to  expansion 
when  the  turbine  is  heated  by  the  admission  of  steam.  The 
connecting  piece  is  used  to  connect  the  turbine  exhaust  and 
the  steam  inlet,  M,  of  the  condenser,  which  may  or  may  not 
be  the  same  shape,  and  also  to  provide  sufficient  head-room 
between  the  turbine  bedplate  and  the  condenser  shell  for  the 
necessary  turbine  supports. 

EXPLANATION. — This  (Fig.  309 A)  gives  the  most  compact  arrangement 
possible  and  requires  a  minimum  of  head-room  and  floor-space.     The 


312  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

condenser  is  placed  directly  beneath  the  turbine,  T,  and  inside  the  tur- 
bine foundation.  All  the  pumps  are  mounted  on  one  shaft  and  driven 
by  one  drive.  The  pump  unit  is  bolted  directly  beneath  the  shell  and  no 
inter-connecting  piping  is  required.  At  the  left  is  the  circulating  pump, 
R;  in  the  center  the  air-pump,  N;  and  at  the  right,  the  condensate  pump, 
P.  The  pumps  may  be  either  directly  driven  by  a  motor,  V,  as  shown  or 
geared  to  a  steam  turbine. 

Ordinarily  the  cooling  water  is  brought  to  the  power  house  through  an 
intake  tunnel,  /,  and  is  discharged  through  a  discharge  tunnel,  D.  The 
water  level  should  be  such  that  the  cooling  water  is  within  the  possible 
suction  lift  for  centrifugal  pumps.  The  suction  piping  should  be  as  short- 
as  possible  to  prevent  air  leaks  and  possible  loss  of  cooling  water  due  to  the 
pump  becoming  air-bound.  If  the  water  levels  permit,  the  discharge 
line,  L,  should  be  brought  back  to  approximately  the  same  level  as  the 
intake  water  and  the  end  of  the  discharge  pipe  sealed  by  extending  it, 
E,  into  the  water.  This  gives  a  siphon  system  (Sec.  374)  and  the  total 
pumping  head  is  then  only  the  friction  head  through  the  piping  and  any 
small  difference  in  the  water  level,  which  may  exist  between  the  suction 
and  discharge  tunnels. 

The  hurling-water  air-pump,  N,  takes  its  hurling  water  from  the 
circulating  pump  discharge  and  discharges,  Q,  to  any  convenient  point 
such  as  the  discharge  tunnel.  The  condensate  pump,  P,  takes  the  con- 
densed steam  from  the  shell,  S,  and  discharges  it  into  the  feed-water 
heater  or  feed  tank.  All  piping — especially  the  circulating  water  piping — 
should  be  made  as  short  as  possible  and  free  of  sharp  bends  which  increase 
the  friction  head.  The  circulating-water  piping  should  have  few  joints 
and  be  free  of  air  leaks  in  order  to  gain  as  much  effect  from  siphonic  action 
as  possible,  and  thereby  maintain  the  circulating-pump  power  at  a 
minimum.  The  piping  should  be  so  arranged  that  no  stress  due  to  expan- 
sion will  be  transmitted  to  the  pump  shell. 

NOTE. — THE  ARRANGEMENT  OF  CONDENSER  PUMPS  SHOWN  IN  Fia. 
309A  Is  USED  FOR  THE  SMALLER  INSTALLATIONS  where  its  compactness 
and  simplicity  make  it  desirable.  For  larger  installations,  separate 
pumps  are  used.  They  may  be  driven  by  one  drive  or  separate  drives 
may  be  provided  for  each  pump.  The  air  and  condensate  pumps 
may  be  combined  and  driven  by  one  drive  and  the  circulating  pump  by 
its  individual  drive.  Motors  or  geared  turbines  are  used  also  for  large 
installations,  the  drive  selected  depending  upon  the  plant  lay-out.  In 
some  cases,  the  circulating  pump  is  driven  by  both  motor  and  turbine 
in  order  to  insure  added  reliability  and  proper  heat  balance.  (See  Sec. 
212). 

356.  The  Principal  Point  To  Be  Observed  In  Caring  For  A 
Condenser  Is  To  Prevent  Leakage  Of  Air.  (H.  H.  Kelley, 
CONDENSERS). — Leaks  may  occur  in  cylinder-heads,  valve- 
chest  covers,  hand-hole  plates,  rod  stuffing  boxes,  flanges  and 


SEC.  357]  STEAM  CONDENSERS  313 

screw  joints  in  piping  (both  exhaust  and  injection  pipes)  and 
around  valve  stems.  Added  to  these  are  the  piston  rod  and 
valve-stem  stuffing  boxes  on  the  engine  and  any  bonnets  that 
may  lie  below  the  exhaust  valves.  The  leaks  are  not  readily 
detected  because  the  pressure  is  on  the  outside  and  the  air  is 
consequently  trying  to  get  in.  A  lighted  candle  or  match 
held  close  to  the  joint  where  the  leak  is  suspected  is  about  the 
simplest  method  of  locating  them.  The  suction  and  discharge 
valves  of  condenser  pumps  should  be  examined  regularly  as 
there  is  no  direct  way  of  detecting  loss  of  vacuum  due  to  poor 
valve  action. 

357.  Strainers  Should  Be  Placed  At  The  Ends  Of  The 
Circulating  Water  Suction  Pipes  whether  the  condenser  takes 
water  directly  from  a  creek  or  pond  or  from  an  intermediate 
reservoir.     Openings  in  the  spraying  device  of  a  jet-condenser 
are  sometimes  comparatively  small  and  these  may  become 
clogged  with  bits  of  foreign  matter  that  might  readily  pass 
through  the  suction  valves.     The  tubes  of  surface  condenses 
may  also  become  clogged  by  foreign  matter  and  thus  decrease 
the  flow  of  circulating  water.     When  a  jet  condenser  fails  to 
get  sufficient  water,  first  examine  the  strainer,  then  the  spray, 
which  can  usually  be  reached  by  removing  the  small  manhole 
plate  on  the  condenser  chamber.     If  trouble  is  not  found  at 
these  points,  examine  the  pump  ports,  the  suction  valves,  dis- 
charge valves  and,  lastly,  the  piston  or  plunger.     If  no  obstruc- 
tion is  found,  the  difficulty  will  be  due  either  to  leakage  of  air 
or  the  heating  of  the  condenser  caused  by  receiving  more  steam 
than  it  is  capable  of  condensing;  in  other  words,  the  condenser 
is  too  small. 

358.  Should  The  Condenser  Vacuum  Suddenly  Decrease 
While  Running,  it  will  probably  be  due  to  an  increased  load 
on  the  engine  and  the   correspondingly  greater  volume  of 
steam   entering   the    condenser.     The    amount    of    injection 
water  which  was  formerly  sufficient  would  then  be  too  small 
for  the  weight  of  the  steam  which  is  to  be  condensed.     The 
obvious  remedy  is  to  open  the  injection  valve.     If  this  does 
not  restore  the  vacuum,  slowly  increase  the  speed  of  the  pump, 
always  watching  the  vacuum  gage,  while  making  these  adjust- 
ments.    If  the  loss  in  vacuum  is  due  merely  to  a  larger  amount 


314  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

of  steam,  these  adjustments  will  restore  it.  If  the  vacuum 
decreases  slowly,  a  little  each  hour  of  the  day,  it  indicates 
leakage  of  air,  a  leaky  piston  and  valves  or  stoppage  of  the 
water  passages  somewhere  between  the  suction  strainer  and 
the  discharge  valves.  The  several  joints  and  the  stuffing  boxes 
may  be  examined  for  air-leaks  in  from  5  to  10  min.  while 
running.  But  an  examination  of  the  valves,  spray  and  pump 
cylinder  can  only  be  made  after  shutting  down  the  condenser. 

NOTE. — IF  THE  CONDENSER  HAS  BECOME  HOT  IT  WILL  NOT  WORK 
UNTIL  IT  Is  COOLED.  As  it  is  necessary  to  bring  steam  in  contact  with 
a  colder  body  in  order  to  condense  it,  should  the  temperature  in  the 
condenser  rise  nearly  to  that  of  atmospheric  exhaust  steam,  condensation 
will  take  place  slowly  and  the  vacuum  can  be  re-attained  only  gradually 
as  the  condenser  cools  again. 

359.  When  The  Atmospheric  Relief  Valve  Of  A  Jet  Con- 
denser Is  Open  And  The  Engine  Is  Running  Non-Condensing, 
Proceed  As  Follows  To  Restore  The  Vacuum  And  Condensing 
Operation. — After  locating  and  removing  any  cause  of  diffi- 
culty, the  pump  may  be  started  and  the  injection  valve  opened. 
The  temperature  will  thus  be  lowered  to  that  of  the  condensing 
water.  With  an  assistant  at  the  atmospheric  relief  valve, 
speed  up  the  pump  and  give  the  condenser  more  water.  Then 
slowly  open  the  stop  valve  in  the  exhaust  pipe,  having  the 
assistant  close  the  relief  valve  at  the  same  rate  as  that  at  which 
the  stop  valve  is  opened.  When  the  relief  valve  is  nearly 
closed,  it  will  close  itself  due  to  the  vacuum  which  will  then 
have  been  produced  in  the  exhaust  pipe,  and  the  engine  will 
run  condensing  again.  The  injection  valve  may  then  be 
partly  closed  and  the  speed  of  the  pump  reduced  a  little,  always 
keeping  watch  of  the  vacuum  gage  while  making  these  ad- 
justments. The  object  is  to  use  as  little  water  as  possible 
and  run  the  pump  as  slow  as  possible  and  still  maintain  the 
desired  vacuum. 

NOTE. — THE  ABOVE  SUGGESTIONS  APPLY  ALSO  To  SURFACE  CONDENS- 
ERS. The  only  difference  is  that  in  some  surface  condenser  plants,  the 
air  pump  and  circulating  pump  are  regulated  separately.  Increasing 
the  speed  of  the  air  pump  is  equivalent  to  increasing  the  speed  of  the 
pump  in  the  jet  condenser.  Increasing  the  speed  of  the  circulating  pump 
has  the  same  effect  as  opening  the  injection  valve  in  the  jet  condenser. 


SEC.  360]  STEAM  CONDENSERS  315 

360.  It  Sometimes  Happens  That  The  Vacuum  Is  Consider- 
ably Below  That  Which  Corresponds  To  The   Condenser 
Temperature,   i.e.,    the   temperature    of  the  condenser  may 
correspond  to  a  vacuum  of  26.5  in.  while  the  highest  vacuum 
which  can  be  maintained  is  25  in.     In  most  instances  this 
will  be  due  to  air  in  the  condenser  and  a  thorough  search  for 
leaks  should  be  made,  provided  the  vacuum  gages  and  ther- 
mometers are  known  to  be  correct.     It  is  practically  impossible 
to   maintain   a   condenser  system   sufficiently  free  from  air 
that. the  vacuum-gage  reading  will  correspond  exactly  to  the 
temperature.     A  reasonable  or  allowable  difference  between 
the  vacuum  gage  reading  and  the  vacuum  corresponding  to 
the  condenser  temperature,   as  found  in  a  steam  table,  is 
about  0.5  in.  mercury  column. 

361.  The  Adjustments  And  Care  Of  The  Barometric  And 
Ejector  Jet  Condensers  consist  largely  of  regulating  the  in- 
jection valve  and  preventing  leaks.     When  a  dry  vacuum 
pump  is  employed  in  connection  with  a  barometric  condenser, 
it  may  need  repair  or  the  speed  may  require  changing  in  case 
of  difficulty  in  maintaining  the  vacuum.     Ordinarily  these 
pumps  are  provided  with  governors,  the  speed  being  changed 
quickly,  when  need  be,  by  adjusting  the  governor. 

362.  With  Surface  Condensers,  Leaky  Tube  Ends  And 
Fouling  Of  The  Tubes  Both  Inside  And  Out  May  Give  Trouble. 
This  condition  shows  itself  in  a  gradually  falling  vacuum. 
Increase  of  the  speed  of  the  air  and  circulating  pumps  affords 
but  temporary  relief.     The  remedy  is  in  thorough  cleaning. 
The  inside  of  the  condenser  may  usually  be  cleaned  with  a 
hose  and  ordinary  city  water  pressure.     A  nozzle  of  pipe 
small  enough  to  go  inside  the  condenser  tubes  is  fitted  to  a 
hose.     A  thick  leather  washer  around  the  nozzle  may  be  used 
to  prevent  the  water  from  squirting  back  and  wetting  the 
operator  when  the  nozzle  is  inserted  in  the  tubes.     If  a  valve 
is  placed  near  the  nozzle,  the  work  may  be  done  by  one  man. 
After  removing  the  head   of  the   condenser,   the   nozzle   is 
pushed  in  and  the  water  is  turned  on.     If  the  water  fails  to 
clean  out  the  tubes,  a  rod  having  a  spiral  end  like  an  auger 
may  be  used  to  scrape  the  tubes  clear  after  which  they  may 
be  rinsed  with  water  as  described  above. 


316  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

363.  When  Grease  Accumulates  On  The  Outside  Of  The 
Condenser  Tubes  it  may  be  removed  by  boiling  the  condenser 
out  with  lye:  Remove  the  handhole  plate  and  put  in  several 
cans  of  lye,  6  or  8  Ib.  for  a  500  h.p.  condenser  and  12  to  15  Ib. 
for  a  1,200  to  2,000  h.p,  condenser.     Provide  a  small  live  steam 
pipe  reaching  well  down  into  the  condenser.    Fill  the  condenser 
with  water.     Heat  the  water  to  the  boiling  point  with  the 
steam  pipe  and  permit  it  to  stand  for  18  to  24  hr.    The  grease 
will  then  run  out  with  the  water — mostly  in  the  form  of  soap. 

364.  An  Index  As  To  The  Condition  Of  Joints  And  Stuffing 
Boxes  Of  Any  Condenser  can  be  obtained  by  noting  the  loss 
of  vacuum  after  shutting  down.     If  all  the  connections,  stuff- 
ing boxes,  and  joints  are  reasonably  tight,  the  loss  of  vacuum 
should  not  exceed  2  in.  per  hr. 

365.  The.  Following  Material  On  Condenser  Selection  And 
Economics  is  based  largely  on  an  article,  APPLICATION  OF 
CONDENSERS,  by  F.  A.  Burg  which  appeared  in  The  Electric 
Journal  for  Dec.,  1920. 

366.  Features  Which  Should  Be  Considered  When  Select- 
ing The  Type  Of  Condenser  To  Use  for  a  given  installation  are 
these:  (1)   The  space  available.     (2)    The  boiler  feed  problem. 
(3)  The  cooling  water.     (4)  Maintenance.     (5)  First  cost.     In 
most  cases,  by  a  general  survey  of  these  items,  the  selection 
can  be  made  without  resorting  to  refinements  and  calculations. 
If,  however,  such  a  survey  shows  that  there  is  little  choice 
between  types,  then  each  type  of  condenser  should  be  con- 
sidered individually.     The  most  economical  size  of  each  type 
should  be  determined,  and  then  these  should  be  compared 
rather  than  arbitrarily  selected.     The  recommended  general 
procedure  in  making  a  selection  is  to  determine  for  each 
condenser  type  under  consideration  the  excess  operating  and 
installing  costs  involved.     Then  when  these  have  been  ascer- 
tained the  propositions  should  be  summarized  and  balanced 
against  one  another.     The  excess  total  annual  operating  and 
maintenance  costs  should  be  capitalized  at  a  reasonable  percent- 
age and  the  resulting  amount  added  to  the  first  cost  of  the 
condenser  that  has  the  excess  operating  cost.     This  total  is  the 
amount  that  it  is  justifiable  to  pay  in  initial  cost  for  the 
condenser  which  effects  the  saving. 


SEC.  367] 


STEAM  CONDENSERS 


317 


EXAMPLE. — Condenser  A  costs  SI, 000  and  its  total  annual  operating 
(power  and  maintenance)  cost  is  $400.  Condenser  B  costs  $700  and  its 
total  annual  operating  cost  is  $500.  Which  of  these  condensers  is  the 
more  economical? 

SOLUTION. — Difference  in  operating  (power  and  maintenance)  cost  = 
$500  —  $400  =  $100  annually.  Assume  a  total  annual  fixed  charge 
(rental  cost  of  space  occupied,  interest,  depreciation,  taxes  and  insurance) 
of  15  per  cent,  on  the  investment.  This  $100  annual  saving  corresponds 
to  a  saving  in  investment  of  $100  -=-0.15  =  $666.70.  Therefore  it  is 
economical  to  pay  $666.70  more  for  condenser  A  than  for  condenser  B. 
But  A  cost  only  $300  more  than  B.  Hence  A  is  the  best  investment. 
Another  method  of  arriving  at  the  same  conclusion  is  to  tabulate  the 
data  thus: 


Item 

A 

First  cost  = 
$1,000 

B 

First  cost  = 
$700 

Operating  cost  

$400 

$500 

Fixed  charge  @  15  per  cent 

150 

105 

Total  annual  charge  

$550 

$605 

Thus  the  data  shows  that  the  yearly  or  annual  cost  of  A  is  $605  -  $550  = 
$55  less  than  that  of  B.  This  $55  annual-cost  saving  would  justify 
an  increase  in  investment  of  $55  -f-  0.15  =  $366.70.  That  is:  $366.70  + 
300  =  $666.70. 

367.  The  Amount  Of  Floor  Space  And  The  Head  Room 
Available  Are  Rarely  Deciding  Factors  In  Selecting  Con- 
densers.— Surface  condensers  require  more  floor  space  than 
do  jet  condensers,  especially  when  allowance  is  made  for  the 
space  required  for  removing  the  tubes.  In  a  new  plant,  space 
for  a  surface  condenser  can  be  provided  without  difficulty,  but 
frequently  turbine  foundations  must  be  specially  designed 
to  accommodate  the  condenser.  The  head  room  required 
For  either  low  level  jet  or  surface  condensers  is  about  the 
same,  if  the  possible  variations  in  design,  such  as  different  shell 
proportions  or  the  use  of  twin  units,  are  recognized.  Generally 
the  question  of  space  is  not  of  primary  importance.  However, 
the  difference  in  the  cost  of  the  installation  due  to  the  differ- 
ence in  space  occupied,  if  any  exists,  should  be  reflected  in 
the  cost  analysis  of  the  problem. 


318  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

368.  The  Quality  Of  The  Available  Feed  Water  Is  Often 
An  Important  Factor  In  Condenser  Selection. — The  surface 
condenser  recovers  the  distilled  condensate  for  boiler  feed 
while  the  jet  does  not.     There   are  relatively  few  natural 
waters  which  do  not  contain  sufficient  solid  matter,  either  in 
suspension  or  solution,  to  form  scale  in  boilers.     Some  waters 
contain  minerals  that  form  a  hard  scale.     Others,  with  just  as 
high  a  mineral  content,  form  a  soft  easily-removable  scale. 
The  questions  of  treating  feed  water,  what  minerals  are  most 
objectionable   and   methods   of   cleaning   boilers   cannot   be 
discussed  here,  but  many  feed  waters  have  to  be  treated. 
The  methods  of  obtaining  good  feed  water  vary  from  a  chemical 
treatment  of  all  of  the  feed  water  to  the  recovery  of  the  condensate 
with  a  surface  condenser,  and  treating  only  the  make-up  water. 

NOTE. — ALTHOUGH  SURFACE  CONDENSERS  SHOULD  DELIVER  PURE 
DISTILLED  WATER  To  THE  FEED  HEATER,  THEY  OFTEN  Do  NOT  Do  So. 
The  purity  of  the  water  depends  on  the  tightness  of  the  tube  packing  and 
the  condition  of  the  tubes  themselves.  If  the  tubes  leak  the  feed  water 
will  be  adulterated  by  the  amount  of  the  leakage.  Hence,  frequent 
electrical  or  chemical  tests  of  the  condensate  should  be  made  to  determine 
its  quality. 

369.  The  Character,  Quantity  And  Source  Of  The  Cooling 
Water  Are  Important  Factors  In  Condenser  Selection. — A  con- 
densing plant  requires  for  condensing  water  alone  from  25  to 
100  Ib.  of  water  per  Ib.  of  steam  condensed.     A  plentiful  supply 
of  water  at  a  low  temperature,   and  at  such  elevation  as  to 
involve  minimum  pumping  power  expense,  is  desirable.     Natu- 
ral heads  are  desirable  but  not  often  available  for  steam  plants. 
Where  the  water  supply  is  limited,  an  artificial  cooling  system 
can  be  installed  (see  Div.  10).     The  amount  of  water  then 
circulated  will  depend  on  the  cooling  range  that  can  be  effected 
by  the  cooling  system  and  not  on  the  type  of  condenser 
employed. 

370.  Cooling  Towers  And  Spray  Ponds  (see  also  Div.  10)  are 
both  used  for  artificial  cooling.     The  rise  in  the  temperature 
of  the  cooling  water  must  be  kept  within  the  cooling  range  of 
the  tower  or  pond,  since  the  water  has  to  be  cooled  in  the  tower 
or  pond  by  the  amount  that  it  has  been  heated  in  passing 
through  the  condenser.     For  the  average  conditions  of  tempera- 


SEC.  371]  STEAM  CONDENSERS  319 

ture  and  humidity,  say  70  deg.  fahr.  air  temperature  and  70 
per  cent,  humidity,  the  cooling  range  for  a  natural-draft  tower 
or  a  spray  pond,  single-spraying,  is  usually  assumed  to  be  from 
14  to  16  deg.  fahr.  and,  for  a  forced-draft  tower  or  a  pond  with 
double-spraying  system,  from  22  to  25  deg.  fahr.  This  means 
that  the  ratio  of  water  to  steam  would  be  between  60  and  70 
to  1  in  the  first  case  and  about  40  to  1  in  the  latter. 

371.  With  Surface  Condensers  Probably  Not  More  Than 
90  To  93  Per  Cent.  Of  The  Boiler  Feed  Will  Be  Returned  To 
The  Boilers.     The  Rest  Will  Have  To  Be  Made  Up.— This 
make-up  water  will,  with  surface  condensers,  have  to  be  treated. 
But  the  expense  of  such  treating  is  small  as  compared  to  the 
expense  of  treating  incurred  with  jet  condensers,  where  all  the 
feed  must  be  treated.     There  will  also  be  a  loss  of  heat  in  the 
feed  when  jet  condensers  are  used  even  if  the  feed  is  taken  from 
the  discharge  of  the  condenser  because  the  temperature  of  the 
condensate  from  a  surface  condenser  is  higher  than  the  tem- 
perature of  the  discharged  cooling  water  from  a  jet  condenser. 

372.  When  Investigating  The  Feed -Water  Phase  Of  The 
Problem  it  will  therefore  be  necessary  to  find  out  the  excess 
cost  of  treating  the  feed,  the  amount  chargeable  to  the  jet 
condenser  for  the  loss  of  heat  in  the  feed  water  and  the  excess 
cost  of  the  treating  plant.     The  cost  of  treating  is  variable.     It 
depends  on  the  nature  of  the  water  to  be  treated.     Ordinarily 
the  cost  does  not  exceed  fifteen  cents  per  thousand  gallons. 
The  loss  of  heat  involved  can  be  reduced  to  the  amount  of 
steam  required  to  raise  the  temperature  of  the  feed  water  to 
that  of  the  condensate  in  a  surface  condenser.     After  this  has 
been  determined  the  cost  of  generating  this  steam  may  be 
ascertained.     The  cost  of  a  treating  plant  will  depend,  on  the 
method  used  and  the  amount  of  water  to  be  treated.     With  all 
these   items   known   another  step  in  the  analysis  has  been 
completed. 

373.  The  Effects  Of  Bad  Water  on  jet  condensers  are  of  less 
moment  than  on  surface  condensers.     In  jet  condensers  the 
parts  subject  to  corrosion  can  be  replaced  more  cheaply.     The 
tubes  in  a  surface  condenser  will  last  indefinitely,  if  the  water 
is  noncorrosive.     But,  surface  condensers  are  frequently  used 
where  only  corrosive  water  is  available.     When  the  water  is 


320  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

quite  bad,  the  tubes  must  be  made  of  a  special  metal  and  even 
then  may  last  only  a  short  time.  When  the  water  is  thus  bad, 
although  it  may  be  highly  desirable  to  save  the  condensate, 
the  cost  of  doing  this  may  not  compare  favorably  with  the  cost 
of  boiler  feed  from  some  other  source. 

374.  The  Most  Important  Phase  Of  The  Cooling  Water 
Problem  Is  The  Cost  Of  Handling  The  Water  under  the  condi- 
tions that  may  exist  in  the  power  plant.     The  jet  condenser, 
by  reason  of  its  ability  to  realize  a  lower  terminal  difference 
(difference  between  the  temperature  of  the  exhaust  steam  and 
that  of  the  outgoing  cooling  water)  does  not  require  as  much 
water  under  average  conditions  as  does  the  surface  condenser. 
This,  however,  does  not  mean  that  it  will  require  less  power. 
With  the  jet  condenser  its  circulating  pump  has  to  pump  all  the 
water  out  of  a  partial  vacuum  which  corresponds  to  about  30 
ft.  head.     In  addition  it  must  discharge  against  an  external 
discharge  head  that  is  never  less  than  the  discharge  head  on 
the  surface   condenser.     The  external  head  consists  of  the 
static  lift  plus  the  friction.     This  means  that  the  jet  condenser 
always  has  a  pumping  head  in  excess  of  thirty  feet,  whereas  the 
surface  condenser  may  not  require  a  head  greater  than  that 
due  to  condenser  and  pipe  friction.     The  head  would  not  be 
greater  than  that  due  to  condenser  and  pipe  friction  where  the 
cooling  water  is  taken  from  a  body  of  water  and   discharged  ^ 
back  at  the  same  level  provided  that  the  whole  system  is  so 
sealed  that  the  full  siphonic  effect  is  realized.     Such  installa- 
tions occur  frequently.     See  Figs.  303  and  309A. 

375.  When  The  Discharge  Level  Is  Higher  Than    The 
Circulating  Pump   (Figs.  310  and  311),   which  condition  is 
ordinarily  encountered  in  spray-pond  installations,  the  advan- 
tage of  lower  pumping  head  is  also  with  the  surface  condenser 
because  the  surface  condenser  can  under  this  arrangement  take 
advantage  of  the  balanced  leg  in  the  circulating  system  while 
the  jet  condenser  cannot. 

EXAMPLE. — Assume  a  cooling-tower  installation  with  the  level  of  the 
cold  well  ten  feet  above  the  circulating  pump.  There  will  be  a  10-ft. 
positive  head  on  the  pump  for  the  surface  condenser  (Fig.  310).  This 
10  ft.  can  be  credited  because,  under  static  conditions,  the  level 
of  the  water  in  the  discharge  pipe  would  be  10  ft.  above  the  pump. 


SEC.  376] 


STEAM  CONDENSERS 


321 


But  a  jet  condenser  (Fig.  311)  cannot  take  advantage  of  this  head  because 
it  would  have  to  pump  the  water  against  a  10-ft.  head  in  addition  to  the 
internal  head  due  to  the  vacuum.  From  this  it  is  evident  that,  in  most 
cases,  the  circulating  pump  of  a  surface  condenser  pumps  against  a  lower 
head  than  does  the  pump  of  an  equivalent  jet  condenser. 


Pump 


FIG.  310. — Showing  Pumping  Head  Of 
Surface-Condenser  Circulating  Pump. 


Fio.  311. — Showing    Pumping    Head 
Jet-Condenser  Circulating  Pump. 


Of 


376.  With  The  Jet  Condenser,  The  Ratio  Of  Water  To 
Steam  Is  Fixed  For  A  Given  Vacuum  Whereas  With  The 
Surface  Condenser  This  Ratio  May  Be  Varied  To  Suit  The 
Conditions. — The  vacuum  obtainable  depends,  with  a  surface 
condenser,  on  a  variety  of  factors,  including  the  rate  of  heat 
transfer  through  the  tubes,  the  velocity  of  the  circulating 
water,  and  the  size  of  the  condenser.  Hence  it  is  possible  to 
select  a  number  of  surface  condensers,  each  with  a  different 
ratio  of  water  circulated  to  cooling  surface,  that  will  produce 
approximately  the  same  vacuum  with  a  given  amount  of 
steam.  But  with  a  jet  condenser,  the  quantity  of  water 
required  is  practically  fixed  when  the  quantity  of  steam  and 
the  vacuum  are  specified.  This  is  due  to  the  fact  that  most 
jet  condensers  realize  a  terminal  difference  in  temperature  of 
between  5  and  8  deg.  fahr.  That  is,  the  rise  in  the  tempera- 
ture of  the  circulating  water  and  the  ratio  of  water  to  steam 
will  be  practically  the  same  for  all  jet  condensers  producing  a 
given  result.  On  the  other  hand,  with  the  surface  condenser, 
the  ratio  of  water  to  steam  may  be  varied  to  suit  the  conditions 
of  different  pumping  heads  and  the  necessity  for  conservation 
of  auxiliary  power. 
21 


322  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

377.  It  Is  Generally  Recognized  That  For  High  Circulating- 
Water  Heads  There  Should,  To  Insure  Minimum  Surface- 
Condenser  Operating  Expense,  Be  A  Lower  Ratio  Of  Water  To 
Surface  Than  For  Low  Heads. — Thus,  where  the  circulating 
water  must  be  pumped  against  a  high  head,  it  is  economical  to 
decrease  the  amount  of  water  to  be  pumped  by  installing  a 
larger  surface  condenser.     That  is,  the  auxiliary  power  and 
therefore  the  operating  expense  may  be  decreased  by  incurring 
a  greater  initial  expense  for  a  larger  condenser.     No  such 
economic  adjustment  is  possible  with  a  jet  condenser.     Since, 
however,  for  a  given  service,  the  jet  condenser  usually  uses 
less  water  than  the  surface  condenser,  it  may  approach  the 
surface  condenser  in  auxiliary  economy  when  the  head  is  high. 
The  handicap  of  the  jet  condenser  circulating  pump  of  having 
to  pump  against  a  greater  head  will  then,  where  the  head  is 
high,  be  offset  by  the  lesser  amount  of  water  to  be  pumped. 

378.  A  Comparison  Of  The  Power  Requirements  Of  The  Jet 
Vs.    The  Surface  Condenser  Should  Not  Be  Based  Solely 
On  A  Consideration  Of  The  Quantities  Of  Water  Circulated 
And  The  Heads  Existing. — There  should  be  considered  also: 
the  facts  that  the  jet  condenser  discharge  pump  is  inherently 
less  efficient  than  a  pump  not  discharging  from  a  vacuum,  and 
that  the  jet  condenser  must  have  a  larger  air  pump  than  the 
surface  condenser.     Against  all  these  jet-condenser  handicaps 
of  less  efficient  pumps  greater  heads  and  more  power  for  the 
air  pumps,  the  jet  condenser  has  the  advantage  of  less  water 
to  circulate.     However,  for  most  installations,  the  jet  con- 
denser requires  more  power  for  drive.     The  amount  of  this 
excess  depends  on  the  discharge  head,  the  type  and  capacity 
of  the  air  pump,  and  the  vacuum  at  which  the  condensers  are 
compared.     With  this  excess  determined,  an  excess  charge 
in  operating  expense  can  be  made.     This  should  be  taken  at  a 
fair  rate  per  horse-power-hour  for  the  total  number  of  hours 
per  year  the  condenser  will  be  in  service.     This  data  provides 
another  item  for  the  final  comparison. 

379.  The  Type  Of  Drive  For  Condenser  Pumps,  Whether 
Electric  or  Steam,  depends  entirely  on  the  use  that  can  be 
made  of  the  exhaust  steam.     If  other  steam-driven  auxili- 
aries, such  as  drives  for  stokers,  fans  and  boiler-feed  pumps, 


SEC.  380]  STEAM  CONDENSERS  323 

furnish  sufficient  exhaust  steam  to  heat  the  feed  water  (Sec. 
265)  it  will  not  be  necessary  nor  economical  to  have  the  con- 
denser auxiliaries  steam  driven.  In  accounting  for  the  excess 
steam  required  by  steam  drives,  it  is  customary  to  disregard  a 
charge  if  all  the  steam  can  be  used  advantageously  in  heating. 
If  the  condenser  auxiliaries  are  motor  driven,  the  charge  is 
usually  determined  by  taking  the  water  rate  on  the  turbine 
from  which  the  motor  derives  its  power,  allowing  for  all 
electrical  losses,  and  thus  arriving  at  the  equivalent  steam 
consumption  per  horse-power  of  the  motor  load.  In  most 
plants  this  will  run  from  fifteen  to  twenty  pounds  of  steam 
per  horse-power-hour.  The  charge  for  the  excess  steam  can 
then  be  determined  from  the  cost  of  producing  the  excess 
steam  that  is  required. 

380.  In  First  Cost  The  Jet  Condenser  Has  A  Decided  Advan- 
tage Over  The  Surface  Condenser. — A  jet  condenser  usually 
costs  about  half  as  much  as  an  equivalent  surface  condenser. 
A  standard  low-level  jet  condenser  for  a  10,000  k.w.  turbine  will, 
at  the  present  prices,  cost  about  $30,000  delivered  and  erected. 
A  surface  condenser  for  the  same  turbine  will  cost  about 
$65,000.     It  is  this  great  difference  in  first  cost  that  often 
renders  the  installation  of  the  jet  condenser  justifiable.     Such 
a  wide  difference  in  first  cost  will  offset  a  considerable  amount 
of  capitalized  savings. 

381.  The  Cost  Of  Maintaining  Pumps  Of  A  Jet  Condenser 
Will  Not  Be  As  High  As  For  The  Surface  Condenser.— This 
is  because  the  jet  condenser  has  only  two  pumps  and  the 
surface  has  three.     But  the  difference  in  these  repair  costs  is  so 
slight  that  it  can  usually  be  neglected. 

NOTE. — PUMP  RUNNERS  MAY  LAST  FROM  A  FEW  MONTHS  To  SEVERAL 
YEARS,  depending  on  the  kind  of  water  being  pumped.  Hence  it  is 
infeasible  to  quote  any  general  data  on  the  cost  of  making  runner 
replacements. 

382.  As  To  The  Relative  Costs  Of  Cleaning  Jet  And  Surface 
Condensers:  the  jet  requires  practically  no  attention,  often 
operating  for  years  without  being  opened.     But  a  surface 
condenser  must  be  cleaned  frequently  to  prevent  a  serious  loss 
of  vacuum.     The  loss  of  vacuum  is  due  to  the  decrease  of  the 


324 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


rate  of  heat  transfer  caused  by  dirty  tubes  and  a  consequent 
restriction  in  the  flow  of  water  through  the  condenser.  The 
frequency  of  cleaning  and  its  cost  vary  with  quality  of  the 
cooling  water.  In  some  plants  condensers  must  be  cleaned 
weekly.  In  others  they  are  cleaned  monthly.  Often  they 


Fio.  311  A. — Worthington  Hydraulic  Tube-Cleaner. 

are  not  cleaned  at  regular  intervals  but  only  after  a  definite 
loss  in  vacuum  has  been  observed.  The  cost  of  cleaning 
varies  with:  (1)  the  character  of  the  deposit  on  the  tube,  (2)  the 
method  of  cleaning,  (3)  facilities  for  handling  the  water  box 
covers,  (4)  the  price  of  labor.  This  cost  may  vary  from  2f£ 
or  3^  per  sq.  ft.  per  yr.  to  15^  or  20^f  per  sq.  ft.  per  yr. 


FIG.   311.B. — Section    Through    Ball-Joint    Of   Worthington    Hydraulic    Tube-Cleaner. 
NOTE. BUILT-IN-TUBE-CLEANING    EQUIPMENT     FOR     SURFACE     CoN- 

DENSERS  (Fig.  311  A)  is  very  useful  and  economical  where  condensers  use 
water  which  is  economical  where  condensers  use  water  which  is  apt  to 
leave  a  deposit  of  vegetable  matter,  mud,  or  slime.  In  the  device  shown 
in  Fig.  311  A,  a  ball  nozzle,  B,  may  be  attached  to  each  manhole  plate  of 
the  condenser.  Through  it  a  cleaning  nozzle,  Nt  may  be  inserted. 
Water  is  led  to  the  nozzle  through  the  hose,  H,  at  a  pressure  of  250  Ib. 
per  sq.  in.  The  nozzle  can  be  swung  to  different  positions  by  handle 


SEC.  383]  STEAM  CONDENSERS  325 

A  on  the  outside  of  the  condenser.  The  nozzle  delivers  about  70  gal.  per 
min.  which,  it  is  claimed,  will  remove  the  mud  from  a  completely-filled 
tube.  With  this  device,  the  condenser  tubes  can  be  cleaned  in  a 
relatively-short  time. 

383.  An  Important  Item  Of  Surf  ace -Condenser    Expense, 
Is  The  Replacement  Of  Tubes. — Tubes  may  last  several  years 
or  they  may  last  only  a  few  months,  depending  on  the  compo- 
sition of  the  tubes  and  the  character  of  the  water.     Instances 
have  been  recorded  where  tubes  have  lasted  for  fifteen  years 
but  this  is  exceptional.     In  industrial  communities,   where 
the  water  is  liable  to  be  contaminated  with  sewage  and  refuse, 
five  or  six  years  is  often  an  average  life.     Here  again  the  limits 
are  so  wide  that  general  figures  would  be  misleading.     At 
the  present  price  of  tubes  assuming  a  5-yr.  life  it  would  cost 
about  ISi  per  sq.  ft.  per  yr.  for  tubes  alone,  disregarding  the 
cost  of  extra  ferrules  and  packings,  and  the  cost  of  installation. 
These  latter  items  would  increase  the  above-quoted  value  to 
at  least  30^f  per  sq.  ft.  per  yr.  for  replacements. 

384.  The  Selection  Of  The  Proper  Condenser  To    Serve 
A  Steam-Driven  Prime  Mover  Is  A  Problem  In  Power -Plant 
Economics.— A  condenser,  regardless  of  type,  is  installed  in  a 
modern  power  plant  only  because  of  the  reduction  which  it 
effects  in  the  cost  of  power.     Operating  condensing,  as  com- 
pared with  non-condensing  operation,  cuts  the  cost  of  power 
approximately  in  half  in  large  turbine  stations.     It  is  therefore 
essential  that  great  care  be  exercised  in  making  the  selection, 
so  that  the  full  saving  from  condensing  operation  may  be 
realized.     Thus  the  problem  of  condenser  application  reduces 
itself  to  a  calculation  to  determine  which  condenser  equipment 
will  produce  power  at  the  least  cost.     The  following  illustrative 
example  explains  the  method : 

EXAMPLE. — Decide  between  a  surface  and  a  low-level  jet  condenser  for 
a  10,000  kw.  plant,  under  the  following  conditions:  200  Ib.  per  sq.  in. 
steam  pressure;  100  deg.  fahr.  super-heat;  space  no  consideration;  boiler 
feed  treating  costs  12£  per  1,000  gal.,  including  chemicals  and  attendance; 
abundant  cooling  water,  available  at  a  head  of  10  ft.,  external  to  the 
condenser;  surface  condenser  maintenance  including  cleaning  and 
replacing  tubes,  35ff  per  sq.  ft.  per  yr.;  maintenance  on  pumps,  the  same  in 
either  case;  plant  operating  7,000  hr.  per  yr.  at  an  average  condensing 
load  of  100,000  Ib.  per  hr. ;  condensers  to  be  selected  on  the  basis  of  75  deg. 


326 


STEAM  POWER  PLANT  AUXILIARIES 


[Div.  9 


fahr.  water  and  28.25  in.  vacuum;  condenser  pumps  to  be  motor  driven 
because  no  exhaust  steam  is  required  for  heating  the  feed  water. 

SOLUTION. — First,  select  the  sizes  of  the  condensers:  The  installation 
will  require  a  jet  condenser  circulating  12,000  gal.  per  min.  or  a  surface 
condenser  having  15,000  sq.  ft.  of  surface  and  circulating  17,500  gal.  per 
min.  to  give  the  required  performance.  The  ratio  of  water  to  surface 
for  the  surface  condenser  has  been  taken  in  accordance  with  common 
practice  for  this  low-head  condition.  The  jet  condenser  will  require 
340  h.p.  for  its  drive  and  the  surface  condenser  will  require  230  h.p. 
The  water  rate  on  the  main  turbine  is  about  10  Ib.  per  b.  hp.h.r.  Assuming 
that  the  motors  will  receive  their  power  from  the  main  unit  at  an  overall 
transmission  efficiency  of  82  per  cent.,  including  generator,  motor,  trans- 
former and  line  losses,  the  steam  per  h.p.  hr.  chargeable  against  the  pumps 
will  be :  10-4-  0.82  or  12.2  Ib.  per  h.p.  hr.  It  has  been  assumed  that  the  cost 
to  generate  the  steam  will  be  35f£  per  l,0001b.  and  the  motor-driven  pumps 
are  charged  on  this  basis.. 

For  this  plant  it  is  assumed,  *if  a  jet  condenser  is  used,  that  the 
water  for  boiler  feed  would  be  taken  from  the  discharge  side  of  the 
condenser  thus  realizing  the  advantage  of  the  higher  temperature. 
When  condensing  100,000  Ib.  of  steam  per  hr.  the  condenser  chosen 
will  have  a  discharge  temperature  of  91  deg.  fahr.  and  thehotwell  tempera- 
ture of  the  surface  condenser  would  be  about  92  deg.  fahr.  There  is 
such  a  slight  difference  in  these  temperatures  that  the  beat  lost  in  the 
feed  when  using  a  jet  condenser  as  compared  that  with  the  surface 
condenser  may  be  disregarded  without  serious  error. 

The  costs  of  operating  the  surface  as  against  the  jet  condenser  may  be 
summarized  thus: 


Items 

Surface 

Jet 

Cost  of  boiler  feed  per  year  
Cost  of  power  for  pumps  
Maintenance,  surface  

$      785 
6,875 
5,250 

$  9,800 
10,160 

Pump  maintenance,  same  

Totals  

$12,910 

$19,960 

Saving  in  favor  of  surface        .      .  . 

$  7,050 

Capitalized  against  jet  @  15% 

47,000 

First  cost  of  condensers  

65,000 

30,000 

Cost  of  water  treating  plant  

2,000 

15,000 

Totals  

$67,000 

$92,000 

From  the  above  tabulation  it  is  evident  that  the  surface  condenser  has  the 
advantage  over  the  jet.  Based  on  the  saving  effected  by  the  surface  con- 
denser, we  could  afford  to  pay  $92,000  for  it  with  the  water  treating 


SEC.  384]  STEAM  CONDENSERS  327 

equipment,  whereas  it  costs  only  $67,  000.  This  same  conclusion  could 
also  have  been  reached  by  calculating  net  savings  instead  of  capitalized 
savings.  But  the  former  method  is  usually  preferable  since  it  indicates 
directly  whether  the  prices  asked  for  the  equipment  are  justifiable. 
The  condensers  used  in  this  comparison  were  selected  on  the  basis  of 
common  practice. 

QUESTIONS  ON  DIVISION  9 

1.  How  does  a  condenser  save  steam?     Increase  power  output? 

2.  Explain    the    operation    of    Newcomen's    condensation    engine.     How   did    Watt 
improve  this  engine? 

3.  What  is  the  function  of  a  condenser  air-pump?     Why  is  it  necessary? 

4.  How  does  the  power  required  by  the  condenser  auxiliaries  compare  with  that 
developed  by  the  condenser? 

How  is  condenser  vacuum  measured?    How  is  it  affected  by  weather  conditions? 
What  is  approximately,  the  most  profitable  vacuum  for  reciprocating  engines? 
For  turbines?     Why  the  difference? 

Give  a  few  advantages  and  disadvantages  of  condensing  operation. 

How  may  condensers  be  classified?     Name  three  classes  of  jet  condensers. 

What  is  the  cooling  medium  commonly  employed  in  condensers? 

10.  Name  three  classes  of  surface  condensers. 

11.  Explain  the  operation  of  a  standard  low-level  jet  condenser  making  a  sketch  of 
the  main  parts.     How  is  water  in  this  condenser  prevented  from  getting  into  the  engine? 

12.  How  is  a  standard  jet  condenser  started  and  stopped? 

13.  How  should  two  jet  condensers  be  connected  when  they  are  used  on  the  same 
exhaust  line? 

14.  How  is  the  air  removed  in  the  ejector  jet  condenser? 

15.  What  are  the  functions  of  the  tail  pipe  of  a  barometric  condenser?     Explain 
how  and  under  what  conditions  a  siphon  or  barometric  condenser  may  be  started  and 
operated  without  a  pump. 

16.  What,  approximately,  should  be  the  velocity  of  the  entering  steam  in  a  jet  con- 
denser?    Velocity  of  cooling  water  issuing  from  a  standard  jet  condenser?     From  an 
ejector  jet  condenser?     From  a  siphon  condenser? 

17.  On  what  does  the  quantity  of  cooling  water  for  a  jet  condenser  depend? 

18.  Explain  by  a  sketch  the  operation  of  a  double-flow,  dry-tube,  horizontal,  surface 
condenser  having  separate  air  and  condensate  pumps.     Explain  the  counterflow  princi- 
ple as  used  in  this  type  of  condenser. 

19.  What  is  the  composition  of  the  tubes,  tube  sheets  and  shells  of  most  surface 
condensers  which  use  salt  water  for  cooling? 

20.  What   factors  determine  the  heat-transfer  coefficient   of  a  surface  condenser? 
Give  approximate  values  for  the  tube  surface  required  per  turbine  kilowatt  developed. 

21.  What  is  meant  by  temperature  "drop"  in  a  condenser?     Give  representative 
values  for  it. 

22.  What  kinds  of  pumps  may  be  used  as  condenser  circulating,  condensate  and  dry- 
air  pumps? 

23.  How  are  reciprocating  pumps  designed  for  high-vacuum  pumping  service? 

24.  Where  may  leaks   occur  so  as  to  impair  condenser  vacuum?     How  may  they  be 
located? 

25.  State  the  usual  causes  of  the  failure  of  a  jet  condenser  to  get  sufficient  water  and 
explain  remedies. 

26.  How    may  an    operator    know  whether  decreased  vacuum  is  due  to  leaks    or  to 
increased  load? 

27.  How  may  a  jet  condenser  be  started  if  it  has  become  hot? 

28.  How  is  a  steam  table  used  in  determining  whether  or  not  a  condenser  is  reasonably 
tight  and  efficient?     What  is  another  way  of  testing  for  tightness? 

29.  Explain  methods  of  cleaning  condenser  tubes  inside  and  outside. 


328  STEAM  POWER  PLANT  AUXILIARIES  [Div.  9 

30.  What  are  the  relative  importance  of  space,  head-room  and   maintenance   changes 
in  selecting  a  condenser? 

31.  How  may  cooling-water  supply  affect  the  selection  of  a  condenser? 

32.  What  is  a  typical  value  for  the  cost  of  purifying  feed  water?     How  does  this 
value  enter  into   condenser  selection?     What  per  cent,  of  the  original  boiler  feed  is 
ordinarily  recovered  by  a  surface  condenser? 

33.  Why  must  a  jet-condenser  circulating  pump  always  work  against  a  30  ft.  greater 
head  than   a   surface-condenser   circulating  pump  under  the  same  conditions?     How 
may  the  pumping  economy  of  a  jet  condenser  equal  that  of  a  surface  condenser  in  spite 
of  this  fact? 

34.  How  do  the  first  costs  of  surface  and  jet  condensers  compare?     Cost  of  cleaning? 

35.  Explain  with  an  example  how  the  economic  advantage  of  two  condensers  may  be 
compared  on  the  basis  of  capitalized  saving. 


PROBLEMS  ON  DIVISION  9 

1.  Steam  is  admitted  to  a  steam  turbine  at  450  deg.  fahr.     It  is  exhausted  at  225  deg. 
fahr.   when  running  non-condensing  and  at  80  deg.   fahr.   when  running  condensing. 
What  are  the  greatest  possible  thermal  efficiencies  when  running  condensing  and  non- 
condensing? 

2.  If  an  engine  has  a  mean  effective  pressure  of  78  Ib.  per  sq.  in.   running   non-con- 
densing, what  will  be  the  saving  in  power  due  to  condensing  operation  with  a  26.5  in. 
vacuum? 

3.  A  turbine  consumes  22  Ib.  of  steam  per  h.p.hr.  at  185  Ib.  per  sq.  in.  abs.  when  run- 
ning non-condensing,  exhausting  against  1  Ib.  per  sq.  in.  back  pressure.     What  will  be 
its  steam  consumption  if  its  exhaust  is  condensed  in  a  29  in.  vacuum? 

4.  The  vacuum  gage  of  a  condenser  indicates  27  in.  of  mercury.     The  barometer 
registers  29.8  in.  of  mercury.     What  is  the  absolute  condenser  pressure  in  inches  of 
mercury?      In  Ib.  per  sq.  in.?     What  per  cent,  of  the  vacuum  possible  at  the  prevailing 
barometric  pressure  does  this  represent? 

5.  A  siphon  jet-condenser  is  required  to  condense  10,000  Ib.  of  exhaust  steam  per 
hour  with  36  Ib.  of  cooling  water  per  pound  of  steam.     The  velocity  of  the  discharge 
through  the  tail-pipe  is  5  ft.  per  sec.     What  should  be  the  volume  of  the  condenser? 
What  should  be  the  diameter,  in  inches,  of  the  tail-pipe? 

6.  The  vacuum  gage  of  a  jet  condenser  registers  27  in.  of  mercury.     The  barometer 
registers  30  in.     The  temperature  of  the  cooling  water  at  entrance  is  80  deg.  fahr.     The 
temperature  of  the  discharge  is  105  deg.  fahr.      The  engine  exhausts  10,000  Ib.  of  steam 
per  hour.     What  is  the  temperature-difference  between  the  discharge-water  and  the 
entering  steam?     How  many  gallons  of  cooling  water  are  required  per  minute? 

7.  The  vacuum  gage  of  a  surface  condenser  registers  28  in.  of  mercury.     The  barome- 
ter registers  29.5  in.     The   condenser  receives   10,000  Ib.  of  exhaust  steam  per  hour. 
The  cooling  water  enters  at  a  temperature  of  67  deg.  fahr.  and  leaves  at  a  temperature 
of  87  deg.  fahr.     The  temperature  of  the  condensate  is  85  deg.  fahr.     How  much  cooling 
water  is  used  per  hour? 

8.  A  surface  condenser  condenses  150,000  Ib.  of  steam  per  hr.  at  an  absolute  condenser 
pressure  of  1.1  in.  of  mercury.     The  circulating  water  enters  the  condenser  at  a  tempera- 
ture of  60  deg.  fahr.  and  leaves  at  a  temperature  of  77  deg.  fahr.    The  condensate  temper- 
ature is  80  deg.  fahr.     The  condenser  is  of  a  modern  dry-tube  type.     The  velocity  of  the 
cooling  water  is  assumed  to  be  about  5  ft.  per  sec.     What  is  the  required  area  of  tube 
surface? 


^ 


DIVISION  10 
METHODS  OF  RECOOLING  CONDENSING  WATER 

385.  Condensing-Water   For  Steam  And  Ammonia  Con- 
densers May  Be  Used  Over  And  Over  Again  if  some  means  for 
re-cooling  it  economically  is  available.     Re-cooling  of  con- 
densing water  may  be  necessary  when,  due  to  material  limita- 
tions, or  for  economic  reasons,  an  ample  supply  of  the  water 
is  unavailable.     Re-cooling  of  the  water  conserves  the  water. 

386.  The  Cooling  Effect  Of  Cooling  Ponds,  Sprays  and 
Cooling  Towers,  on  condensing  water  is  due  to  three  causes. 
(1)  Evaporation  of  the  water.     (2)  Direct  heat  transfer  by  con- 
duction and  convection,   which   is  of  minor   consequence   as 
compared  to  that  of  the  evaporative  effect.     (3)  Direct  heat 
transfer  by  radiation  which  also  is  of  minor  consequence.     The 
cooling  effect  of  evaporation  is  due  (See  the  author's  PRACTICAL 
HEAT)  to  the  fact  that  whenever  a  liquid  evaporates — when 
it  is  transformed  into  a  vapor — an  amount  of  heat  equivalent 
to  its  latent  heat  of  vaporization  must  be  absorbed  by  it  to 
effect    the    vaporization.     In    the    atmospheric    cooling    of 
condensing  water,  practically  all  of  this  heat  which  is  required 
to  effect  the  evaporation  of  the  condensing  water  is  abstracted 
from  the  unvaporized  portion  of  the  condensing  water  itself. 
Thereby  the  remaining  portion  of  the  water  is  cooled.     A 
minor  portion  of  this  heat  which  is  required  to  effect  the 
evaporation  is  abstracted  from  adjacent  air  and  objects. 

NOTE. — The  cooling  effect  of  the  direct  heat-transfer  (See  the  author's 
PRACTICAL  HEAT)  is  usually  of  minor  consequence;  see  Note  under 
Sec.  399.  This  direct-heat-transfer  cooling  effect  is  caused  by  the  heat 
in  the  condensing  water  being  conducted  and  radiated  into  the  surround- 
ing air  and  objects. 

387.  Atmospheric  Recooling  Of  Condensing-Water,  after 
the  water  has  been  discharged  from  the  condensers,  may  be 
promoted  by  bringing  the  water  into  intimate  contact  with 
the  air  of  the  atmosphere  and  by  the  evaporation  of  a  part 

32? 


330 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


of  the  water.  Intimacy  of  contact  with  the  air  and  ample 
surface  to  promote  effective  evaporation  is  secured  by  break- 
ing up  the  mass  of  water  into  a  fine  spray  or  into  a  multitude 
of  tiny  streams  or  rivulets  or  by  spreading  it  out  over  an 
extensive  area  in  a  shallow  pond.  The  recooling  effect 
depends  upon:  (1)  The  temperature-difference  between  the 
water  and  the  air.  (2)  The  relative  humidity  of  the  air.  High 
humidity  and  high  air-temperatures  are  drawbacks  to  satis- 
factory re-cooling.  (3)  The  degree  of  contact-intimacy. 


Distributer-    L~iacfafcr 

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Ner 
-y 


;-Distr!bufer    I 


Cooling- 
Tower 


•By- Pass 


Cast -Iron 
Deflector- 


—  Louvres  — 


Collecting 
•Pan 


Atmospheric-, 
Condenser  ' 


Motor-Driven  Centrifugal: 
Circulating  Pump •' 

FIQ.  312. — "Burhorn"  Metallic  Tower 
For  Cooling  The  Water  For  Double- 
Pipe  Ammonia  Condenser. 


FIG.  313. — "Burhorn"  Metallic  Tower 
For  Cooling  The  Water  For  An  Atmos- 
pheric Ammonia  Condenser. 


E. — Profitable  operation  of  an  atmospheric  recooling  system  is, 
in  general,  mainly  dependent  upon  the  degree  of  effectiveness  with  which 
all  parts  of  the  water  are  brought  into  contact  with  the  air. 

NOTES. — MODERN-TYPE  STEAM-CONDENSERS  ORDINARILY  OPERATE 
(Sec.  328)  WITH  VACUA  RANGING  FROM  ABOUT  26  IN.  To  28  IN.  of 
mercury  column.  Assuming  the  temperature  of  the  water  entering 
the  condenser  to  be  about  85  deg.  fahr.,  the  discharge  temperature, 
corresponding  to  the  vacua  above  noted,  would  range  from  about  90  to 
110  deg.  fahr. 


SEC.  388]  METHODS  OF  RECOOLING  CONDENSING  WATER  331 


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332 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


THE  TEMPERATURE  OF  THE  WATER  LEAVING  AN  AMMONIA  CONDENSER 
(Figs.  312  and  313)  of  the  submerged  type  may  be  from  about  75  to 
80  deg.  fahr.,  while  the  water  from  an  atmospheric  ammonia-condenser 
may  during  the  summer  months  have  a  temperature  of  from  about  75  to 
85  deg.  fahr.  This  subject  is  discussed  in  the  Author's  MECHANICAL 
REFRIGERATION. 

THE  TEMPERATURE  AND  RELATIVE  HUMIDITY  OF  THE  ATMOSPHERIC 
AIR  (Table  388)  are  dependent  upon  the  locality  and  the  season  of  the 
year.  For  practical  purposes,  the  local  weather  bureau  reports  may  be 
referred  to  for  this  information.  But  where  these  are  not  obtainable, 
the  relative  humidity  (see  the  author's  PRACTICAL  HEAT)  must  be  deter- 
mined (Sec.  389)  by  the  use  of  instruments  made  for  the  purpose. 

389.  To  Determine  The  Relative  Humidity  Of  The  Air,  a 

sling  psychrometer  (Fig.  314)  may  be  used.  This  instrument  is 
formed  with  two  ordinary  thermometers.  The  dry  bulb  of  one 
Tij  is  dry  and  bare,  so  as  to  be  exposed 
directly  to  the  temperature  of  the  air. 
The  wet  bulb  of  the  other,  T2,  is  covered 
with  cotton  gauze  or  cloth  which  is  satu- 
rated with  water. 

EXPLANATION. — If  air  is  blown  over  the  two 
thermometers,  or  if  they  are  swung  by  rotating 
the  handle,  H,  rapidly  through  the  air,  the  one 
having  the  "wet  bulb"  will  generally  show  a 
lower  reading  than  the  one  with  the  "dry  bulb." 
This  is  due  to  the  fact  that,  in  general,  atmos- 
pheric air  is  not  fully  saturated  (see  the  author's 
PRACTICAL  HEAT).  It  will  still  have  some 
capacity  for  absorption  of  moisture.  Therefore 
it  will  absorb  moisture  from  the  wet  gauze  which 
envelopes  the  "wet  bulb."  A  cooling  effect,  due 
to  evaporation  of  the  moisture  in  the  gauze,  is 
thereby  produced. 


Supporting  Eye 


FIG.  314. —  Sling  Psy- 
chrometer For  Determin- 
ing Relative  Humidity. 


390.  The  Relative  Humidity  Of  The 
Air  Is  A  Function  Of  The  Temperature  - 
Difference  Indicated  By  The  Wet-  And  Dry-Bulb  Ther- 
mometers Of  A  Sling  Psychrometer  (Fig.  314).  Hence,  when 
the  temperature-difference  shown  thereby  is  known,  the 
corresponding  relative  humidity  may  be  computed  therefrom, 
or  it  may  be  obtained  directly  from  the  results  of  such  com- 
putations, which  are  given  in  Table  393.  How  these  relative- 


SEC.  391]  METHODS  OF  RECOOLING  CONDENSING  WATER  333 

humidity  values  are  utilized  in  practical  computations  will  be 
hereinafter  explained. 

NOTE. — When  there  is  no  difference  (Table  393),  between  the  readings 
of  the  wet-  and  dry-bulb  thermometers  (Fig.  314),  then  the  air  is  fully 
saturated  with  moisture.  That  is  (Sec.  387),  the  air  has  absorbed  as 
much  water  as  it  can  possibly  retain,  at  the  given  temperature,  in  a 
vaporous  condition.  Hence,  no  cooling  effect,  due  to  evaporation  from 
the  wet  bulb,  can  result.  The  relative  humidity  is  then  100  per  cent, 
(see  the  author's  PRACTICAL  HEAT). 

EXAMPLE. — When  the  dry-bulb  thermometer  (Fig.  314)  reads  70  deg. 
fahr.,  and  the  wet-bulb  thermometer  reads  60  deg.  fahr.,  the  temperature- 
difference  =  70  —  60  =  10  deg.  fahr.  The  corresponding  relative  humid- 
ity, from  Table  393,  is  55  per  cent.  This  value  is  found  in  the  same 
horizontal  column  with  the  given  value,  70,  of  the  air-temperature  and 
in  the  same  vertical  column  with  the  computed  value,  10,  of  the  tempera- 
ture-difference. 

l/391.  The  Limit  Of  Atmospheric  Cooling  Is  The  Wet-Bulb 
Thermometer  Temperature. — Careful  investigation  proves 
that  this  is  the  lowest  temperature  attainable  by  cooling  in 
fee  contact  with  the  atmosphere  (COOLING  TOWER  COMPANY). 
lThis  temperature  is,  then,  a  measure  of  the  efficiency  of  any 
atmospheric-cooling  device.  Perfect  apparatus,  that  having 
an  efficiency  of  100  per  cent,  would  reduce  the  temperature 
of  the  cooled  water  to  that  of  the  wet  bulb,  The  number  of 
degrees  temperature  decrease  thus  effected,  would  be  the 
ideal  range.  The  number  of  degrees  temperature  decrease 
attained  in  practice  is  the  actual  range.  Hence:  Actual  range 
-f-  Ideal  range  =  Efficiency  of  the  apparatus,  or  the  percentage 
of  the  ideal  which  is  actually  realized.  See  Sec.  392  for  the 
formula  which  expresses  this  relation. 

NOTE.— The  wet-bulb  temperature,  therefore,  bears  the  same  relation 
to  atmospheric  cooling  that  the  barometic  height  does  to  condenser 
vacua.  It  is  the  ideal  minimum  temperature  which  can  be  approached 
infinitely  close  but  which  can  never  be  passed.  How  nearly  this  ideal 
minimum  temperature  may  be  attained  is  determined  by:  (1)  Water  dis- 
tribution. (2)  Cooling  surface.  (3)  Air  supply.  Increasing  the  effec- 
tiveness of  any  or  all  of  these  elements  decreases  the:  first  cost,  operating 
expense,  and  maintenance  expense.  There  is  then,  a  certain  degree  of 
attainment  toward  the  ideal  past  which  it  does  not  pay — in  dollars  and 


334  STEAM  POWER  PLANT  AUXILIARIES         [Drv.  10 

cents — to  proceed.  The  determination  of  this  "point  of  maximum  econ- 
omic effectiveness"  is  a  problem  for  specialists. 

392.  The  Efficiency  Of  Any  Atmospheric  Cooling  Device, 

cooling  pond,  spray  nozzle  installation  or  cooling  tower,  may 
be  computed  by  the  following  formula : 

(92)  E  =  100  -wP^sF  (per  cent.) 

4/1/T    1  fw 

Wherein  E  =  the  efficiency,  in  per  cent.  T/i  =  the  tempera- 
ture, in  degrees  Fahrenheit,  of  the  water  coming  to  the  cooling 
device.  T7/2  =  the  temperature,  in  degrees  Fahrenheit,  of 
the  cooled  water  leaving  the  device.  T/w  =  the  wet -bulb 
temperature  of  the  surrounding  atmosphere  in  degrees  Fahren- 
heit, corresponding  to  the  given  relative  humidity,  as  com- 
puted from  Table  393. 

EXAMPLE. — The  temperature  of  the  water  entering  a  cooling-tower  is 
108  deg.  fahr.  The  temperature  of  the  water  leaving  the  tower  is  88  deg. 
fahr.  The  temperature  and  relative  humidity  of  ithe  outside  air  are, 
respectively,  70  deg.  fahr.  and  50  per  cent.  What  is  the  efficiency  of 
the  tower? 

SOLUTION. — By  Table  393,  the  difference  between  a  dry-bulb  tempera- 
ture of  70  deg.  fahr.  and  the  corresponding  wet-bulb  temperature,  for 
51  per  cent,  relative  humidity,  is  11  deg.  fahr.,  while  the  difference  for 
48  per  cent,  relative  humidity  is  12  deg.  fahr.  Therefore,  the  wet-bulb 
temperature  corresponding  to  50  per  cent,  relative  humidity  =  70  —  jll  + 
[(12  -  11)  -r-  (51  -48)1}  =  68.7  deg.  fahr.  Then,  by  For.  (92),  the 
efficiency  of  the  tower  =  E  =  100[(77/1  -  Tf2)/(Tfl  -  Tfv}}}  =  100  X 
[(108  -  88)  -=-  (108  -  68.7)]  =  51  percent. 


SEC.  393]  METHODS  OF  RECOOLING  CONDENSING  WATER  335 


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336 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


.012 


394.  The  Weight  Of  Water  Vapor  Which  Is  Contained  In 
A  Cubic  Foot  Of  Atmospheric  Air  Is  Determined  By  The  Tem- 
perature And  The  Relative  Humidity  Of  The  Air. — The  graph 
Fig.  315  indicates  the  relation  between  temperature  and  weight 

for  air  of  100  per  cent,  relative  hu- 
midity. To  obtain  the  weight  at 
any  relative  humidity  other  than 
100  per  cent.,  multiply  the  value 
taken  from  the  graph  by  the 
known  relative  humidity  expressed 

ahrenheit      decimally. 


9      2     40    60    80    100  12 
Air  Temperature  in  Degree 


FIG.  315.—  Graph  Showing  Re-        EXAMPLE.  —  The  temperature  of  a  cer- 


Ft.  Of  Air  Of  100  Per  Cent.  Rela-  relative  humidity  is  55  per  cent.  What 
tive  Humidity.  is  the  weight,  per  cubic  foot,  of  its 

moisture     content.      SOLUTION.  —  From 

the  graph  of  Fig.  315,  the  weight  of  the  water-vapor  content  in  1  cu. 
ft.  of  air,  at  100  deg.  fahr.  and  at  100  per  cent,  relative  humidity,  is  0.003 
Ib.  Hence,  for  55  per  cent,  relative  humidity,  the  moisture  content  of 
the  air  =  0.003  X  0.55  =  0.00165  Ib.  per  cu.  ft. 

NOTE.  —  When  air  is  "saturated,"  its  relative  humidity  is  then  100 
per  cent,  and  the  weight  of  water-vapor  content  in  it  is  a  maximum. 
Hence,  for  saturated  air,  the  weight  of  its  water-vapor  content  is  deter- 
mined solely  by  the  temperature.     Like- 
wise,    assuming     any    constant    relative 
humidity,  the  weight  of  the  water  vapor 
content  will  be  determined  solely  by  the 
temperature. 


395.  The  Water  Vapor  Pressure 
Exerted  By  Water  Vapor  In  Air  Is 
Determined  By  Its  Temperature 
And  By  The  Relative  Humidity.— 
Water-vapor-pressure  values  are 


Temp,  of  Air  in  Degrees 


FIG.  316.  —  Graph  Showing  Re- 
lation Between  The  Temperature 
And  Vapor  Pressure  Of  Saturated 
Water  Vapor  (Or  Of  Water  Vapor 

used  in  computing  the  effectiveness    in  Air  Of  100  Per  Cent.  HU- 


of    cooling  ponds  and  towers  and    midity)-    These  A™  M^eiy  vai- 

.      ..  ues  Plotted  From  A  Steam  Table. 

similar     condensing-  water  -cooling 

arrangements.  The  graph  of  Fig.  316  shows  the  relation 
between  temperature  and  vapor  pressure  for  saturated-air 
water  vapor,  that  is,  for  the  water  vapor  in  air  which  is  of  100 
per  cent,  humidity. 


SEC.  396]  METHODS  OF  RECOOLING  CONDENSING  WATER  337 

NOTE. — To  OBTAIN  THE  WATER- VAPOR  PRESSURE  EXERTED  BY 
VAPOR  IN  UN-SATURATED  AIR,  multiply  the  pressure  value  (taken  from 
Fig.  316)  which  corresponds  to  the  known  temperature,  by  the  relative 
humidity,  expressed  decimally. 

396.  Three  Principal  Devices  For  Bringing  The  Water  And 
Air  Into  Intimate  Contact  In  A  Recooling  System  are  com- 
monly available.  These  are:  (1)  The  simple  cooling  pond  or 
tank  (Fig.  317).  (2)  The  spray-fountain  (Fig.  318).  (3) 
The  cooling-tower  (Fig.  319).  Each  of  these  devices  has  its 
particular  field  of  application,  as  will  be  shown  in  following 
Sees. 


' .' •.  :'.:•  •'• '.' V  •' :.':'Cono/er?sing '-tfafer 

':  Shallow  Trough  Exposing  Lcrye  Wafer? Area.±  '••  '.• ':'.';  .'•'••! ''••:  ••  :' 


.--•WctTer  Should  Discharge  -into  Lna  o1 
Pond  Opposite  from  Intake  I 


FIG.  317. — Diagrammatic    View   Of  A  Typical   Cooling  Pond.      (A  ditch  may  be  sub- 
stituted for  the  trough  T). 

397.  Cooling  Ponds  May  Satisfy  The  Requirements  Of  A 
Recooling  System  Where  Ample  Ground  Space  Is  Available. 

The  operation-expense  of  a  simple  cooling  pond  is  very  low. 
The  power-cost  may  be,  and  often  is,  practically  zero.  Gen- 
erally, however,  for  plants  exceeding  about  1,000-h.p.  capacity, 
the  area  and  investment  necessary  for  an  adequate  cooling 
pond  would  be  so  extensive  that  the  annual  cost  of  the  pond 
would  be  prohibition.  Hence,  for  the  larger  plants,  more 
compact  devices,  as  spray  fountains  and  cooling-towers,  may 
be  more  economical  and  satisfactory. 

398.  The  Rate  Of  Evaporation  From  A  Simple  Cooling 
Pond,  When  The  Air  Is  Perfectly  Calm,  may  be  computed 
by  the  following  formula : 

(93)  w  =  (243  +  3.7T/)(P,  -  PVM)  (  grains  per  sq.  ft.  perhr.) 

22 


338 


STEAM  POWER  PLANT  AUXILIARIES          [Div.  10 


Wherein  W  =  the  weight  of  water  evaporated,  under  calm 
air,  in  grains  per  square  foot  per  hour;  it  may  be  increased 
materially  by  the  effect  of  wind;  see  following  note.  T/  = 
the  temperature  of  the  water,  in  degrees  Fahrenheit.  Pv  = 


TtoTor-  Driren 
Cenfrifugal. 
Spray  Pump. 


Flo.  318. — 'Diagram  Showing  Schutte  And  Koerting  Double  Spraying  System.  In 
Winter  One  Side  Is  Shut  Down.  (Spray  nozzles  are  set  at  an  angle  of  45  degrees  to 
the  horizontal.) 

the  vapor  pressure,  in  inches  of  mercury-column,  as  taken 
from  the  graph  (Fig.  316),  for  saturated  air  at  the  given 
temperature.  M  =  the  per  cent.,  expressed  decimally,  of 
the  relative  humidity  of  the  air,  as  found  in  Table  393. 


SEC.  398]  METHODS  OF  RECOOLING  CONDENSING  WATER  339 

NOTE. — In  the  expression  "  (Pv  —  PVM}"  in  the  above  equation,  "Pv" 
is  the  vapor  pressure  which  would  be  exerted  by  a  saturated  air  vapor 
at  the  given  temperature  and  "PVM"  is  the  vapor  pressure  actually 
exerted  by  the  -non-saturated  air  vapor  under  consideration.  Their 
difference  is  a  measure  of  the  tendency  to  promote  evaporation.  See 
Sec.  395. 


Tile 
(Filling 


tton  Pipe     .-  '••  LOWer  oyerf  low  Pipe'Closed  "       ''••  "••  '  Lower  Overf  low 
I-Cold-Well  Level  Raised  To  Permit  I-CoId-Well  Water  Leval  Lowered  For 

Forced  Draft  Operation  Natural  Draft  Operation 

FIG.  319.  —  Worthington  Cooling-Tower  Using  Either  Forced  Or  Natural  Draft. 


NOTE.  —  IN  PRACTICE,  THE  WEIGHT  OF  THE  EVAPORATION  may,  due 
to  normal  wind  velocities,  be  from  2  to  12  times  greater  than  the  value 
obtained  by  the  preceding  formula.  A  fair  average  is  from  6  to  8  times 
the  computed  value. 

NOTE.  —  For.  (93)  may  be  rewritten  as  follows: 

(243  +  3.7Tf)(Pv  -PVM) 


(94) 


W.  = 


7000 


(Ib.  per  sq.  ft.  perhr.) 


Wherein  Ww  =  weight  of  water  evaporated,  under  calm  air,  in  pounds 
per  square  foot  per  hour.     T/,  Pv  and  M  are  as  given  in  For.  (93). 


340  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

EXAMPLE. — The  temperature  of  the  water  in  a  cooling  pond  is  80  deg. 
fahr.  The  air-temperature  is  also  80  deg.  fahr.  The  relative  humidity 
is  75  per  cent.  Assuming  that  the  prevailing  wind-velocity  multiplies, 
8  times,  the  rate  of  evaporation  under  calm  air  conditions,  what  is  the 
approximate  rate  of  evaporation,  in  pounds  per  square  foot  per  hour? 
How  many  square  feet  of  the  pond  surface  are  required  to  give  off  each 
pound  of  the  evaporation?  SOLUTION. — The  graph  of  Fig.  316  shows 
the  water  vapor  presswe  corresponding  to  a  temperature  of  80  deg.  fahr.  = 
1.0  in.  of  mercury-column.  By  For.  (93),  the  rate  of  evaporation  in  calm 
air  =  W  =  (243  +  3.7Tf)(Pv  -  PVM)  =  (243  +  3.7  X  80)  X  [1  -  (1  X 
0.75)]  =  134.75  grains  per  sq.  ft.  per  hr.  The  avoirdupois  pound  con- 
tains 7,000  grains.  For  the  prevailing  wind-velocity,  the  approximate 
rate  of  evaporation  =  (134.75  X  8)  -=-  7,000  =  0.154  Ib.  per  sq.  ft.  per  hr. 
Therefore,  1  -f-  0.154  =  6.49  sq.  ft.  per  Ib.  per  hr.  =  number  of  square 
feet  of  pond  surface  necessary  to  evaporate  1  Ib.  of  water  per  hour. 

399.  The  Evaporation  Of  One  Pound  (1  Ib.)  Of  Condensing 
Water  Is  Equivalent  To  The  Abstraction  Of  About  1,000  B.t.u. 
From  The  Water. — This  is  true  because,  as  shown  in  any 
steam  table,  the  latent  heat  of  vaporization  (or  evaporation) 
of  water  vapor  is,  for   the  vapor  pressures  encountered  in 
cooling-pond,  spray-nozzle  and  cooling-tower  practice,  about 
1,000  B.t.u.     The  approximate  vapor  pressure  in  any  instance 
is  that,  as  taken  from  the  graph  of  Fig.  316  corresponding  to 
the  existing  air  temperature.     The  exact  vapor  pressure  is 
that  taken  from  Fig.  316  for  the  given  air  temperature,  multi- 
plied by  the  relative  humidity.     The  relative  humidity  may 
be  obtained  as  explained  in  Sec.  390. 

EXAMPLE. — If  8  Ib.  of  water  evaporates  from  the  water  in  a  cooling 
pond,  a  spray  pond  or  a  cooling  tower,  there  will  thereby  be  abstracted 
from  the  water  in  the  pond  approximately:  8  X  1,000  =  8,000  B.t.u. 

NOTE. — With  low  air-temperatures,  radiation  of  heat  from  the  pond, 
and  transfer  of  heat  to  the  air  by  conduction  and  convection,  assist,  to 
some  extent,  in  cooling  the  water.  The  loss  of  heat  by  evaporation  may 
under  this  condition,  be  somewhat  reduced.  With  high  air-temperatures, 
the  reverse  is  true.  It  is  generally  observed  that,  in  moderately -warm 
weather  and  under  ordinary  conditions,  approximately  90  per  cent,  of 
the  cooling  effect  is  due  to  evaporation  and  10  per  cent,  to  other  causes. 

400.  In  Estimating  The  Requisite   Surface  Area  For  A 
Simple  Cooling  Pond,  for  cooling  the  circulating  water  of  a 
steam  condenser,  (Fig.  320)  it  may,  safely,  be  assumed  that: 
(1)  The  total  heat  given  up  or  lost  by  the  cooling-pond  water  is 


SEC.  400]  METHODS  OF  RECOOLING  CONDENSING  WATER  341 

solely  that  abstracted  from  the  water  by  evaporation.  (2)  The 
total  heat  imparted  to  the  water  is  the  heat  given  thereto  in  the 
condenser  by  the  steam  during  its  condensation  therein.  Hence, 
if  the  temperature  of  the  condensing  water  in  the  pond  is  to 
be  maintained  constant,  the  pond  area  must  be  sufficiently 
great  that  it  will,  by  its  evaporative  effect,  release  the  same 
amount  of  heat  per  hour  as  is  imparted  to  it  per  hour  by  the 
condensing  steam.  Now,  it  may  also  be  assumed  that  the 


-=>         -"^>  "ft    Engine,         Make-Up 
Surface    r  L^sL*"""  *•*      Wbtfw- 


I-With    Jet   Condenser  (Circulating 


IE-  Wit  h  JetCondenser  (Circulating  Wafer  Unsuitable,  for  Bofter  Feed) 
FIG.  320. — Three  Possible  Arrangements  Of  Condensing  Equipment. 

heat,  in  B.t.u.,  which  is  given  up  to  the  condensing  water  by 
1  Ib.  of  steam  when  the  steam  is  condensed,  is  equal  to  the 
heat,  in  B.t.u.,  which  is  abstracted  from  the  cooling-pond 
water  by  1  Ib.  of  the  water  when  it  evaporates.  In  both  cases, 
the  amount  of  heat  is  about  1,000  B.t.u.: 

Therefore:  The  approximate  requisite  total  pond-area  will 
result  if  the  area  (Sec.  398)  which  is  required  to  give  off  1  Ib.  of 
evaporation  per  hour,  is  multiplied  by  the  number  of  pounds 
of  steam  which  is  condensed  per  hour. 

NOTE. — THE  ABOVE  ASSUMPTIONS  ARE  NOT  STRICTLY  ACCURATE. 
But  inasmuch  as  the  resulting  values  which. are  obtained  by  using  For. 


342  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

(93)  and  (94)  must  be  increased  (Sec.  398)  by  from  2  to  12  times  to  cor- 
rect for  wind  effect,  the  above-indicated  method  is  sufficiently  accurate 
for  estimating. 

NOTE. — THE  COOLING  EFFECT  OF  THE  "MAKE-UP"  WATER  MAY  BE 
DISREGARDED  because  the  make-up  water — that  which  must  be  replen- 
ished because  of  evaporization,  windage  and  other  losses — is  less  than 
2  to  3  per  cent,  of  the  total  amount  of  water  which  is  circulated.  The 
variation  in  evaporative  effect  due  to  wind  (Sec.  398)  will  more  than 
offset  the  cooling  effect  of  the  make-up  water. 

EXAMPLE. — A  steam  condenser  (any  type)  is  required  to  condense 
3,600  Ib.  of  exhaust  steam  per  hour.  The  water  will  be  discharged  to  a 
cooling  pond  for  recooling.  The  temperature  of  the  discharge  water 
will  be  80  deg.  fahr.  The  air  temperature  is  also  80  deg.  fahr.  The 
relative  humidity  is  75  per  cent.  What  should  be  the  area,  in  square 
feet,  of  the  cooling  pond? 

SOLUTION. — It  is  computed  in  the  Example  under  Sec.  398  that,  under 
the  conditions  just  specified,  each  6.5  sq.  ft.  of  pond  area  will  evaporate 
1  Ib.  of  water  per  hour.  Hence  (Sec.  400:  Total  pond  area  =  Area  re- 
quired to  give  of  1  Ib.  evaporation  per  hour  X  Number  of  pounds  steam 
condensed  per  hour),  the  total  pond  area  should  be:  6.5  X  3,600  =  23,400 
sq.  ft. 

401.  The   Requisite   Area  For   A   Simple   Cooling  Pond 
Cannot   In   Any  Case  Be  Computed  Precisely,  due  to  the 
numerous  variables  which  are  involved.     The  most  important 
of  these  are  (1)  the  temperature  and  humidity  of  the  air  (Sec. 
387)  (2)  the  solar  reheating  effect  and,  particularly  (3)  the  wind- 
velocity  (Sec.  398).     The  type  of  condenser  and  the  kind  of 
condenser-service,    whether   steam   or   ammonia,    may    also 
affect  the  problem. 

NOTE. — WHEN  A  COOLING  POND  Is  LOCATED  ON  THE  ROOF  OF  A 
BUILDING,  THERE  Is  MORE  SOLAR  REHEATING.  Hence,  in  such  loca- 
tions, the  pond  area  must,  usually,  be  greater  for  an  equivalent  effect. 
If  spray  nozzles  are  used  over  a  roof  pond,  the  same  number  spaced  in 
the  same  way  will  not,  ordinarily,  give  as  good  results  as  over  a  surface 
pond. 

402.  Some    Cooling -Pond -Area    Data    Are:  It    has  been 
determined  (PROCEEDINGS  A.  S.  M.  E.,  Apr.,  1912,  page  607) 
that,  in  the  northern  part  of  the  United  States,  120  sq.  ft.  of 
cooling-pond  surface  will  suffice  for  1  h.p.  of  steam-condenser 
service.     This  value  is  based  upon  a  26-in.  vacuum  and  a 
steam  consumption  of  15  Ib.  per  h.p.  per  hr.     Cooling-pond 
area  is  sometimes  determined  upon  the  assumption  that  8  sq. 


SEC.  403]  METHODS  OF  RECOOLING  CONDENSING  WATER  343 

ft.  will  suffice  for  each  pound  of  steam  condensed.  Also, 
that  1  sq.  ft.  of  pond  area  will  give  off  4  B.t.u.  per  hour  per 
deg.  fahr.  difference  between  the  water-and  air-  temperature 
in  summer,  and  2  B.t.u.  in  winter. 

403.  The  Depth  Of  Simple  Cooling-Ponds  is  usually  from 
3  to  4  ft.     The  depth  has  little  influence  on  cooling  effective- 
ness, provided  the  surface-area  is  ample,  since  the  cooling  is 
determined  almost  wholly  by  the  surface  area  which  is  exposed 
to  the  air  and  from  which  evaporation  can  take  place. 

404.  Spray    Fountains    Are    Often    Used    In    Connection 
With  Cooling-Ponds.— This  arrangement  (Fig.  318  and  321) 


MM 


Spray        Pond     '• 


:  Pipes  :- 


Concrete  Pier-, 


Discharge 
PipeFrom 
Condenser 


^  . 


Fia.  321.  —  -Spray  Pond  With  Cooling-Tower  Company's  "Impact"  Nozzles.     (Space 
between  sprays  permits  effective  air  circulation.) 

permits  of  a  considerable  reduction  in  the  area  of  the  pond. 
The  fine  division  of  the  water  particles  by  the  sprays  (Fig. 
318)  insures  a  maximum  of  water  surface  in  small  space  and 
thereby  facilitates  the  cooling  effect  due  to  evaporation  and 
air  contact. 


NOTE. — The  passage  of  the  water  through  the  cores  of  the  nozzles 
(Figs.  322  and  323)  on  a  spray  fountain,  breaks  it  (Figs.  324,  325  and  326) 
into  a  fine  mist.  These  nozzles  are,  generally,  set  either  vertically 
(Fig.  324),  at  an  angle  of  45  deg.  with  the  horizontal  (Fig.  318),  or  at  an 


344 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


angle  of  60  deg.  with  the  horizontal  (J,  Fig.  327).  These  arrangements 
secure  a  wide  distribution  of  the  spray.  They  also  tend  to  induce  air- 
currents,  even  on  calm  days,  which  greatly  augment  the  cooling  effect. 


•Outlet 


Stotno/oirct  Pipe  Thread,  Usuvllyl$To  3  //7.--> 


FlG.    32  2. — Badger    Spray    Nozzle. 
(Badger  &  Sons  Co.,  Boston.) 


< -Inner 

Nozzle 


Body-. 


f?-Sforno/vrol  Pipe  Thread,  Usually  1$  To  5  In. 
FIG.    323.— Koerting  Multi-Spray  Nozzle. 


FIG.  324. — Form  Of  Spray  From  Single  Spray-Nozzle. 

405.  The  Conditions  Which  Mainly  Control  The  Amount 
Of  Recooling  Produced  By  Spray  Fountains  have  been  deter- 
mined by  tests.  It  has  been  demonstrated  (Fig.  328) :  (1) 


SEC.  405]  METHODS  OF  RECOOLING  CONDENSING  WATER  345 

That   recooling   is   more   affected   by   the   air-temperature   and 
humidity  than  by  the  temperature  of  the  water  coming  from  the 


^^^^HB||iiP%. 


FIG.  325.— Form  Of  Intermingled  Spray  From  Three  Nozzles. 

condensers.  (2)  That  with  80  to  90  per  cent,  relative  humidity, 
the  water-temperature  can  be  lowered  to  within  12  or  13  deg. 
fahr.  of  the  dry-bulb  air-temperature.  (3)  That  with  20  to  30 


^Xx\\\\l! '///;-/, 

V\f;oi  n-Shoi  peoT.-v. 
XSpriu^/^^^- 


fr7fersec//n^l-j^;-rj-^|  Connection 
Jets      '    ' 


Fia.  326. — Cooling  Tower  Company's  Impact  Spray  Nozzle.     (Designed  to  minimize 
possibility  of  clogging  and  to  promote  air  circulation.) 


per  cent,  relative  humidity,  the  water-temperature  can  be  lowered 
about  8  deg.  fahr.  below  the  dry-bulb  air-temperature.  (4)  That 
the  loss  of  water  is  usually  about  6  per  cent. 


346 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


FIG.  327.— Utilizing  Roof-Space  For  Spray-Cooling.     (Koerting.) 


«     8    10    12    14    16 


56   40    42    U    4« 


aterOn 


v— Cloudy 


v-C/owcrt/  Cloudy—-*  *'Rct!rt         I 

Clear  Except  As  Noted  On  The  W,10"Anot  //*-•' 


Fia.  328. — Graph  Of  Spray-Nozzle  Tests  On  A  "  Cooling-Tower-Company  "  Installa- 
tion At  Silver  Springs,  New  York,  September,  1919.  (Flat  spray  nozzles  were  used. 
Circulation  =  5,000  gal.  per  min.  on  steam-condenser  duty.  The  capacity  (size)  of 
these  nozzles  was  60  gal.  per  min.  at  6M  lb.  per  sq.  in.  pressure.  The  value  of  "  Kn" 
used  in  the  guarantee  equation,  For.  (95),  was  "5.7.") 


SEC.  406]  METHODS  OF  RECOOLING  CONDENSING  WATER  347 

406.  To  Compute  The  Temperature  Reduction  Which  Can 
Be  Effected  By  A  Spray-Nozzle  Installation  the  following 
formula  can  be  used.  It  is  quoted  from  THE  COOLING  TOWER 
COMPANY'S  CATALOGUE  and  is  the  basis  of  its  guarantees. 


(95)  T/2  = 


+460)  +  (TV, +  460)  1  •  _  j£  + 

2  J 


Kn  X  100,000,000 


Wherein,  all  temperatures  are  in  degrees  Fahrenheit  and; 
T/z  =  temperature  of  cooled  water  after  spraying.  T r/i  = 
temperature  of  water  before  spraying.  T /x  =  (4TT/W  +  T/d) 
-f-  5.  T fd  =  dry-bulb-thermometer  or  air  temperature.  T/w 
=  wet-bulb-thermometer  temperature.  Kn  =  a  constant  = 
5.1  for  average  installations  operating  at  6J£  lb  water  pres- 
sure but  it  may  vary  from  4.0  to  5.7.  These  values  for  Kn 
were  determined  from  tests  made  by  the  COOLING-TOWER 
COMPANY  using  the  impact  nozzle  of  Fig.  326.  Kn  varies 
with  the  type,  size  and  spacing  of  the  nozzles  and  with  the 
water  pressure  and  wind  velocity.  For  equal  operating  pres- 
sures and  atmospheric  conditions,  the  value  of  Kn  depends 
mainly  on  the  pond  exposure,  the  size  of  the  nozzles  and  the 
ratio  of  pond  area  to  water  sprayed. 

NOTE. — THE  PREDETERMINATION  OP  THE  PROPER  VALUE  FOR  Kn, 
for  any  given  installation,  requires  extensive  experience  in  this  branch 
of  engineering.  Consequently,  to  design  a  cooling  system  which  will 
develop  a  given  value  of  Kn,  a  thorough  knowledge  of  the  local  conditions 
is  necessary  as  well  as  a  practical  understanding  as  to  the  effects  of  such 
conditions.  It  is  feasible,  should  the  service  conditions  justify  the  expen- 
diture, to  so  design  the  system  that  the  value  of  Kn  will  be  as  low  as  4.0 
or  even  less. 

EXAMPLE. — See  Fig.  328  which  indicates  the  approximate  agreement  of 
of  actual  observed  values  with  values  obtained  by  computation  with 
For.  (95)  using  a  value  of  5.7  for  Kn. 

NOTE. — By  using  values  from  Table  388,  the  probable  temperature 
reduction  which  may  be  expected  in  any  locality  can  be  computed. 

NOTE. — PERFORMANCE  GUARANTEES  ON  COMBINED  CONDENSER-AND- 
SpRAY-CooLiNG  OUTFITS  can  be  obtained  from  certain  manufacturers — 
Schutte  &  Koerting  Co.  for  example.  In  such  guarantees,  the  vacuum 
performance  of  the  condenser  is  based  on  the  outside-air  temperature — 


348 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


not  on  the  temperature  of  the  injection  water;  a  standard  relative  humid- 
ity as  is  assumed. 

407.  The  Size  And  Number  Of  Nozzles  To  Be  Used  In  A 
Spray-Fountain  (Table  408)  depends  upon  the  quantity  of 
water  to  be  handled.     It  is  commonly  assumed  that  a  single 
spraying  system  will,  under  normal  conditions,  cool  the  water 
about   20   to    30   deg.   Fahr.     This   is   considered   sufficient 
(Table  410)  for  ordinary  steam-condenser  service.     However, 
it  is  often  considered  desirable  to  spray  from  25  to  55  per  cent, 
of   the   condensing  water  a  second  time  before  sending  it 
through  the  condenser. 

408.  Table  Showing  Spray-Nozzle  Capacities  In  Gallons 
Per  Minute.     (SCHTJTTE  &  KOERTING  COMPANY). 

NOTE. — Nozzles  of  2-in.  pipe-size  are  most  frequently  used.  These  are 
commonly  regarded  as  the  most  economical.  The  outlet  orifice  in  the 
tip  of  a  2-in.  nozzle  has  a  diameter  of  about  0.8  in.  The  hydraulic 
pressure  required  to  force  the  water  through  the  nozzles  should  never 
exceed  about  14  Ib.  per  sq.  in.,  gage. 


Pipe-size  of 
nozzle,  in  inches 

Pressures  on  nozzles,  in  pounds  per  square  inch 

5 

6 

7 

8 

9 

10 

2 

54 

60 

65.5 

70.5 

75 

78 

2M 

77 

85 

92 

98 

103 

106 

3 

115 

125 

133 

140 

146 

151 

409.  The  Spacing  Of  The  Nozzles  In  A  Spray -Fountain 
depends  mainly  upon  the  design  and  size  of  the  nozzles. 
Centrifugal  nozzles  of  2-in.  size  are  usually  spaced  about  8  to 
10  ft.  from  center  to  center.  Nozzles  of  larger  size  may  be 
set  proportionately  further  apart. 

NOTE. — A  typical  installation,  spraying  4,800  gal.  per  min.,  consists  of 
9  rows  of  nozzles,  with  8  nozzles  in  each  row.  Thus,  each  nozzle  sprays 
4,800  -=-(9X8)=  66?3  gal.  per  min.  The  rows  are  20  ft.  apart,  and  the 
nozzles  are  spaced  13  ft.  between  centers.  A  2-in.  nozzle  (Table  408)  at 
a  little  over  7  Ib.  per  sq.  in.  water  pressure  would  meet  these  requirements. 


SEC.  410]  METHODS  OF  RECOOLING  CONDENSING  WATER  349 


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350  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

411.  The  Ground -Area  Required  For  Spray -Fountain  Ponds 

is  much  less  than  that  required  for  simple  cooling-ponds.  It  is 
commonly  assumed  that  for  small  plants,  under  500  h.  p.,  1  sq. 
ft.  of  surface  will  suffice  for  the  cooling  of  150  Ib.  of  water  per 
hr.  For  plants  of  about  5,000  h.p.,  1  sq.  ft.  of  surface  will 
usually  suffice  for  the  cooling  of  250  Ib.  of  water  per  hr.  For 
plants  of  about  1,000  h.  p.,  1  sq.  ft.  of  surface  may  be  assumed 
as  sufficient  for  the  cooling  of  200  Ib.  of  water  per  hr. 

NOTE. — SPRAY-FOUNTAINS  SHOULD  BE  SURROUNDED  BY  WIND- 
BREAKS, in  order  to  avoid  excessive  water-loss,  due  to  heavy  winds. 

412.  Spray-Fountains  Are  Sometimes  Located  On  Power- 
House  Roofs  (Fig.  327).     This  is  usually  done  where  ground- 
area  is  unavailable.     The  extra  power  required  for  elevating 
the  condensing  water  may,  with  some  types  of  condenser 
installations  (Fig.  327),  be  offset  by  utilizing  the  hydrostatic 
head,  thus  obtained,  for  sending  the  recooled  water  through 
the  condenser. 

EXPLANATION. — The  water  from  the  hot-well  (Fig.  327)  is  pumped  to 
the  spray-nozzles  J,  by  the  centrifugal  pump,  G,  through  the  discharge- 
pipe,  H.  The  recooled  water  in  the  spray,  V,  runs  into  the  trough  K. 
It  then  flows  through  the  low-level-jet  eductor  condenser  C,  under  the 
head  due  to  its  elevation  above  the  hot-well. 

413.  The  Power  Required  To  Operate  A  Spray-Fountain 

in  connection  with  a  steam-condenser  equipment  is  not  great. 
The  pressures  generally  used  seldom  require  more  than  1.5  to 
2  per  cent,  of  the  main-engine  power.  This  is  equivalent  to 
about  10  per  cent,  of  the  power  saved  by  condensing  over  non- 
condensing  operation.  Very  often,  the  circulating  pump  (Sec. 
353)  may  be  used  to  deliver  the  water  directly  to  the  nozzles. 
Installation  of  additional  pumping  equipment  is  thereby 
avoided.  This  can  generally  be  done  with  surface-condensers 
(Sec.  335)  and  low-level  jet-  or  eductor-condensers  (Sec.  336) 
but  not  (Sec.  339)  with  barometric  condensers. 

414.  A  Cooling -Tower  (Fig.  319)  consists,  essentially,  of  a 
tall,   narrow,   wooden   or   sheet-iron   structure,   so   arranged 
internally  that  after  the  warm  condensing-water  has  been 
elevated  to  the  top  under  pump-pressure,  it  will  fall,  by  gravity, 
in  a  multitude  of  thin  sheets  or  trickling  streams,  into  a  reser- 


SEC.  414]  METHODS  OF  RECOOLING  CONDENSING  WATER  351 
voir  or  sump,  S,  which  is  located  beneath  the  tower.     In  falling 


Notches : 


Cypress  ; 
Plonks--"' 


Water  Supplied  Here- 


Cypress  Trough 


•Wafer  Sf reams-  > 
.-Initial  Spray- 


*•— Nozzles— -*' 


.--Galvanized 
'  Iron  Troughs-, 


i     Secondary  Breaking  Up  ,     Cypress  Triangular 
^  \  of  Water 'Particles-.    ,-,          BaffMs.^ 

:%7'      \>    '"^571      V'  *'^j^>:       V    ;"^p,'l. 

1       '       i\7  .        \     '      '  "^^         .      '<       l'^^~         I     '\      ("*tfP\ 


r^;  u        >     *      • 


&> 


FIG.     329.— Cypress    Board    Checker        FIG.  330. — ''Wheeler"  Cooling-Tower  Splash 
Work  For  Cooling  Towers.  Counter-Flow  System. 


-Swamp  Cedar  Lumber 


I-     Plan       view    Of    Deck 


,-NozzIe 


.-Distributer 


Cast-Iron  Gutter-^ 


Swamp  Cedar--''     '.9 

sinter  mediate  Deck    *J 


Distributing  Deck-\' 

(I !' 


I-Beo/m-'' 


-  Section  Through  Distributor  hnd  Decks 


FIG.  331.  —  Distributor  And  Decks  Of  "The  Cooling-Tower  Company"  Design.     (Each 
tower  contains  10  intermediate  decks  arranged  one  above  the  other.) 

it  is  cooled  by  the  air  which  surrounds  it.     The  devices  for 
dividing  the  water  into  fine  sheets,  droplets  or  sprays  may 


352 


STEAM  POWER  PLANT  AUXILIARIES          [Div.  10 


consist  of:  (1)  Checker  work  (Fig.  329).  (2)  Corrugated  sur- 
faces. (3)  Troughs  or  baffles  (Fig.  330  and  331) .  (4)  Galvanized 
steel  wire  screens  or  perforated  trays.  (5)  Masses  of  tile-tubing 
or  galvanized  iron  pipes  (Fig.  332)  set  vertically. 


^-Interlocking  P/pes... 


c 


^•Air  And  Wafer-- 
Spaces 


FIG.  332. — Interlocking   Pipe   Filling   In    Mixing   Chamber   Of   Worthington    Tower. 

NOTE. — In  every  case,  the  tower  is  open  at  the  top,  and  is  so  arranged 
at  the  bottom  that  atmospheric-air  will  circulate  (either  by  natural  draft 
or  by  pressure  of  a  fan-blower)  through  the  descending  water.  The 
water  gives  up  its  heat  to  the  ascending  air-currents  by  evaporation, 
convection  and  radiation. 

Pertorvfect  Pipe  Distributer.;. 
Trough  Deck- 


^'•Collecting  Pan 

Fia.  333.— "  Burhorn "  Open  Or  Atmospheric  Cooling  Tower.  (Louvres  removed 
from  one  side  to  show  construction.  In  some  cities  wooden  cooling  towers  are  pro- 
hibited because  of  fire  risk.) 

NOTE. — FROM  75  To  85  PER  CENT.  OF  THE  RECOOLING  EFFECTED 
IN  A  COOLING-TOWER  RESULTS  FROM  EVAPORATION  in  most  power 
plant  installations.  The  percentage  of  recooling  effected  by  conduction, 


SEC.  415]  METHODS  OF  RECOOLING  CONDENSING  WATER  353 


connection  and  radiation,  both  in  the  tower  and  from  the  piping  which 
conveys  the  water  thereto,  is  usually,  in  power  plants,  comparatively 
insignificant  probably  rarely  exceeding  more  than  2  per  cent.  But  where 
towers  are  used  for  cooling  water  from  high  temperature  stills,  where  the 
cooling  range  may  be  100  deg.  fahr.  or  more,  then,  the  combined  cooling 
effect  due  to  radiation  and  conduction  may  be  greater  than  that  due  to 
evaporation. 


Viypor  . 
Outlet--' 


FIG.  334. — Section  Of  Typical  Wheeler-        FIG.     335. — Worthington     Forced    Draft 


Balcke  Natural  Draft  Cooling-Tower. 


Cooling-Tower. 


WOOD  CHECKER  WORK  (Fig.  329)  FOR  COOLING  TOWERS  usually  con- 
sists of  1  X  4-in.  cypress  or  swamp-cedar  boards  set  on  edge  and  spaced 
about  4  in.  apart. 

416.  Cooling  Towers  May  Be  Divided  Into  Four  General 
Classes:  (1)  Open  or  atmospheric-towers  (Fig.  333)  using 
natural  draft.  (2}  Closed  or  chimney-flue  towers  (Fig.  334) 

23 


354 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


using  natural  draft.     (3)  Closed  or  chimney-flue  towers  (Figs. 
335  and  336)   using  forced  draft.     (4)  Closed  or  flue-towers 


Steam  Pipe-. 
•Generator       .'Turbine 


FIG.  336.— Forced    Draft    Cooling    Tower    With    Surface    Condenser.      (Worthington 

Company.) 

(Fig.  319)  using  either  forced  or  natural  draft.  The  sides  of 
open  or  atmospheric  towers  (Fig.  333)  are  usually  louvred, 
(Fig.  337)  to  prevent  the  water  from  being  blown  out  of  the 


!•  Isometric     View 
Fia.  337. — "Burhorn"  Sheet-Metal  Cooling-Tower  Louvres. 

tower.     Louvres  actually  decrease  the  cooling  effect  but  must 
be  employed  to  minimize  water  waste. 


SEC.  416]  METHODS  OF  RECOOLING  CONDENSING  WATER  355 

416.  The  Closed  Or  Flue-Towers  Are  Completely  Enclosed, 
Except  At  Top  And  Bottom. — Openings  are  provided  in  the 
base  for  admission  of  the  fan  blast  in  the  one  case  or  the 
natural   air-currents  in  the  other.     Natural  draft  in  these 
towers  depends  entirely  upon  the  chimney  action  of  the  tower. 

NOTE. — CHOICE  OF  FORCED  OR  NATURAL  DRAFT  mainly  depends  up- 
on space  considerations  on  the  one  hand  and  operating-cost  on  the  other. 
A  forced-draft  installation  occupies  less  space  than  one  using  natural 
draft,  but  the  operating  expense  is  greater.  Where  cooling-towers  are 
designed  (Fig.  319)  for  using  either  forced  or  natural  draft,  the  forced 
draft  may  be  used  during  the  hot  season  and  the  natural  draft  in  cool 
weather. 

THE  OPEN  TOWER  (Fig.  333)  PERMITS  A  SOMEWHAT  GREATER  Loss 
OF  WATER  THAN  THE  CLOSED  TOWER  (Fig.  334).  This  is  due  to  winds 
blowing  through  the  louvres  of  the  open  tower.  Generally,  the  air  does 
not  mingle  so  effectively  with  the  water  in'  open  towers  as  it  does  in 
closed  towers,  but  the  closed  tower  must  be  larger  for  the  same  cooling 
effect.  In  many  fan  (forced-draft)  towers,  the  water  lost  is  greater  than 
in  an  atmospheric  tower  of  about  the  same  size.  This  is  because  a  large 
amount  of  water,  as  entrained  moisture,  is  carried  away  in  the  forced- 
draft  towers  due  to  the  high  air  velocity.  Since  such  water  has  not  been 
evaporated,  it  represents  pure  waste — it  has  performed  no  useful  work 
of  cooling  by  its  evaporation.  With  a  forced-draft  and  an  atmospheric 
tower  operating  side  by  side,  the  water  loss  from  the  forced-draft  tower 
may  be  as  great  as  10  per  cent,  and  that  from  the  atmospheric  tower  as 
small  as  2  per  cent. 

417.  The  Principles  Involved  In  Cooling-Tower  Computa- 
tions are  similar  to  those  pertaining  to  cooling-ponds  and 
spray-fountains.     The  cooling  effect  depends  upon  the  water- 
and  air-temperatures,  the  relative  humidity  of  the  air,  and  the 
effectiveness  of  air-and-water  contact.     Towers  of  different 
types  vary  in  the  effectiveness  with  which  the  air  is  utilized 
as  a  cooling  medium. 

418.  Computations    To    Determine    Cooling-Tower    Per- 
formance Should  Be  Based  On  The  Results  Of  Tests  And 
Practice  rather  than  on  entirely  theoretical  assumptions.     If 
the  condition  of  the  atmosphere  as  to  temperature  and  humi- 
dity, the  temperature  of  the  water  coming  from  the  condensers, 
the  quantity  of  water  each  unit-volume  of  air  will  absorb, 

/  and  the  degree  of  efficiency  under  which  the  tower  will  operate, 
known,   then  reasonably-close   approximations  may  be 


356  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

made  for  any  specific  case  by  applying  the  general  methods 
of  computation  (Sec.  398)  previously  given  for  cooling-ponds. 
The  general  method  is  illustrated  in  a  following  example. 

NOTE. — In  the  operation  of  a  cooling  tower,  the  same  water  is  used 
over  and  over  again.  Through  the  process  of  cooling  there  is  a  certain 
loss  which  must  be  made  up  from  some  outside  source.  The  water  which 
must  be  supplied  to  compensate  for  this  loss  is  known  as  the  make-up 
water.  Make-up  water  is  equal  to:  (water  lost  by  evaporation)  +  (water 
which  is  splashed  or  blown  out  of  the  cooling  lower.)  Assuming  that  there 
is  no  loss  except  that  due  to  evaporation,  the  amount  of  heat  (in  B.t.u.) 
taken  away  from  the  water  in  circulation,  will  (See  Sec.  400)  equal  the 
number  of  pounds  of  water  lost,  multiplied  by  approximately  1,000. 
In  other  words,  every  pound  of  water  evaporated  will  carry  away  1,000 
B.t.u.,  and  cool  1,000  Ib.  of  water  1  deg.  fahr.,  or  100  Ib.  of  water  10  deg. 
fahr.,  etc.  Therefore,  to  cool  100  Ib.  of  water  10  deg  fahr.,  requires  the 
evaporation  of  1  Ib.  of  water,  or  1  per  cent,  of  the  amount  cooled.  Thus, 
theoretically,  the  make-up  water  will  be  1  per  cent,  of  the  water  circu- 
lated, to  cool  the  water  10  degrees.  Actual  tests  on  several  Burhorn 
towers  under  different  conditions,  have  shown  the  actual  loss  to  be 
less  than  1%  per  cent,  of  the  total  amount  circulated,  or  practically 
that  due  to  evaporation. 

Under  usual  ammonia-condenser  conditions,  a  cooling  tower  may  be 
expected  to  cool  the  water  by  from  5  to  14  deg.  fahr. ;  about  10  deg.  fahr. 
is  a  reasonable  expectancy.  For  steam  condensers,  a  tower  may  be 
expected  to  decrease  the  temperature  of  the  water  by  from  20  to  50  deg. 
fahr. 

419.  To  Compute  The  Average  Temperature  Reduction 
Effected  In  Summer  Weather  By  Atmospheric  Cooling 
Towers  now  in  operation  in  this  country  and  abroad,  use  the 
following  empirical  formula  which  is  derived  from  the  results 
of  a  large  number  of  tests.  It  is  quoted  from  THE  COOLING 
TOWER  COMPANY'S  CATALOGUE. 

(96)  Tf.  =  Tfd  +  2l"°  +  Tn  (deg.  fahr.) 

Wherein,  all  temperatures  are  in  degrees  Fahrenheit  and: — 
T /a  =  average  temperature,  of  the  cooled  water  which  leaves 
cooling  towers.  Tfd  =  dry-bulb-thermometer  or  air  temper- 
ature. Tfw  =  wet-bulb-thermometer  temperature.  T/i  = 
temperature  of  water  entering  the  cooling  tower. 


SEC.  420]  METHODS  OF  RECOOLING  CONDENSING  WATER  357 

EXAMPLES. — See  Tables  421  and  422  which  show  average  values  com- 
puted with  the  preceding  formula.  By  using  values  from  Table  388, 
the  probable  temperature  reduction  which  may  be  expected  in  any 
locality  can  be  computed. 

NOTE. — Cooling  towers  can  be  designed  which  will,  for  certain  cooling 
ranges  and  atmospheric  conditions,  reduce  the  cooled-water  temperature 
by  from  10  to  50  per  cent,  below  that  given  by  the  preceding  formula. 
The  possible  maximum  temperature  reduction  is  determined  by  the 
cooling  range  and  by  atmospheric  conditions.  See  Tables  421  and  422. 

420.  Typical  Data  Pertaining  To  Cooling-Tower  Perform- 
ance have  been  obtained  from  a  series  of  tests  made  with 
closed  cooling-towers  using  forced  draft.  They  are  as  follows : 

DATA. — Quantity  of  water  circulated  =  640  gal.  per  min.  Tempera- 
ture of  air  entering  the  tower  =  70  deg.  fahr.  Temperature  of  air 
leaving  the  tower  =  94  deg.  fahr.  Relative  humidity  of  air  entering 
the  tower  =  50  per  cent.  Relative  humidity  of  air  leaving  the  tower  = 
100  per  cent.  Temperature  of  water  entering  the  tower  =  108  deg. 
fahr.  Temperature  of  water  leaving  the  tower  =  88  deg.  fahr.  quantity 
of  air  circulated  =  50,000  cu.  ft.  per  min.  Efficiency  of  tower  (For. 
92)  =  51  per  cent.  These  data  represent  about  average  practice  for 
the  given  type  of  installation. 

EXAMPLE. — Using  the  above  data,  and  allowing  8.3  Ib.  to  the  gallon, 
the  heat  added  to  the  water  while  passing  through  the  condenser  = 
640  X  8.3  X  (108  -  88)  =  106,240  B.t.u.  per  min.  Assuming  the 
specific  heat  of  air  to  be  0.019  B.t.u.  per  cu.  ft.,  the  heat  which  the  air  ab- 
sorbs, by  convection  and  radiation,  from  the  water  in  the  tower  =  50,000  X 
0.019  X  (94  -  70)  =  22,800  B.t.u.  per  min.  =  (22,800  -=-  106,240)  X 
100  =  21.46  per  cent,  of  the  heat  which  the  water  absorbed  in  the  con- 
denser. Hence,  the  heat  which  the  water  gives  off  by  evaporation  = 
106,240  -  22,800  =  83,440  B.t.u.  per  min.  =  (83,440  -5-  106,240)  X 
100  =  78.54  per  cent,  of  the  heat  which  the  water  absorbed  in  the  con- 
denser. Assuming  (Sec.  400)  that  each  pound  of  the  evaporation  ab- 
stracts 1,000  B.t.u.,  the  water-loss  =  83,440  -5-  1,000  =  83.44  Ib.  per 
min.  =  83.44  -r-  (640  X  8.3)  X  100  =  1.57  per  cent.  Wind  losses  might 
increase  this  to  over  2  per  cent.  In  practice,  the  usual  water  loss  may  be 
from  2  to  3  per  cent. 

NOTE. — THE  PER  CENT.  OF  WATER-LOSS  FROM  COOLING  TOWERS,  as 
noted  above,  is  less  than  the  lowest  per  cent,  of  loss  that  can  be  obtained 
with  spray-fountains.  This  is  an  important  item  in  favor  of  the  cooling- 
tower. 


358 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 


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SEC.  422]  METHODS  OF  RECOOLING  CONDENSING  WATER  359 


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360  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

423.  A  Method  Of  Computing  The  Proportions  Of  A  Cool- 
ing-Tower will  now  be  explained  by  the  use  of  an  illustrative 
example.  Cooling-tower  design  is — because  of  the  necessity 
of  using  results  from  existing  installations  as  precedents — 
properly  a  function  of  men  of  considerable  experience  in  this 
particular  branch  of  engineering. 

v  EXAMPLE. — A  forced-draft  cooling-tower  is  required  to  re-cool  1,000,000 
Ib.  of  condensing  water  per  hour  through  25  deg.  fahr.  The  circulating 
air  is  assumed  to  be  at  a  temperature  of  75  deg.  fahr.  when  it  enters  the 
tower  and  at  105  deg.  fahr.  when  it  leaves.  What  should  be:  (1)  The 
total  horizontal  cross-seciional  area?  (2)  The  total  horizontal  length  of  each 
tide?  (3)  The  total  height  of  the  checkerwork? 

yr  SOLUTION. — It  may  be  assumed  that  the  tower  is  to  be  furnished  with 
a  cypress-board  checker-work  (Fig.  329)  for  dividing  the  descending  water 
into  a  multitude  of  thin  sheets.  Practice  has  shown  that  air  velocities, 
in  cooling-towers,  of  about  700  ft.  per  min.  produce  the  best  results. 
It  may  be  assumed  that  the  evaporating  or  cooling  surface  afforded  by 
the  cypress  boards  is  about  8  sq.  ft.  per  cubic  foot  of  space  occupied  by 
the  checkerwork.  It  may  further  be  assumed  that  about  64  per  cent, 
of  the  total  horizontal  cross-sectional  area  of  the  checkerwork  is  effective 
area,  or  free  area.  Also  that  20  B.t.u.  per  hour,  per  degree  of  cooling, 
will  be  abstracted  from  the  condensing  water  for  each  square  foot  of 
evaporating  surface. 

The  water  will  absorb,  in  the  condenser  approximately  1  B.t.u.  per  Ib. 
for  each  deg.  fahr.  of  temperature  increase.  Hence,  the  total  quantity  of 
heat  to  be  abstracted  in  the  cooling-tower  =  25  X  1,000,000  =  25,000,000 
B.t.u.  per  hour.  The  quantity  of  heat  abstracted  per  square  foot  of  cypress- 
board  evaporating  surface  =  20  X  25  =  500  B.t.u.  per  hour.  Therefore, 
the  requisite  total  area  of  evaporating  surface  =  25,000,000  -r-  500  = 
50,000  sq.  ft.  Hence,  the  total  volume  of  space  to  be  occupied  by  the  checker- 
wood  =  50,000  -f-  8  =  6,250  cu.  ft. 

Assuming  (Example  subjoined  to  Sec.  420)  that  about  21.5  per  cent.  = 
0.215  of  the  heat  in  the  condensing  water  passes  to  the  air  by  convection 
and  radiation,  the  total  quantity  of  heat  so  removed  =  25,000,000  X  0.215 
=  5,375,000  B.t.u.  per  hour.  Therefore,  assuming  the  specific  heat  of 
air  to  be  0.019  B.t.u.  per  cu.  ft.,  the  requisite  quantity  of  air,  of  the  given 
entering  and  leaving  temperature,  =  5,375,000  -5-  [0.019  X  (105  -  75) 
=  9.429,824  cu.  ft.  per  hr.  =  9,429,824  -s-  60  =  157,167  cu.  ft.  per  min. 

For  an  air-velocity  of  700  ft.  per  min.,  the  requisite  effective  cross- 
sectional  area  of  the  checker  work  =  157,167  -r-  700  =  224.5  sq.  ft. 
This  being  about  64  per  cent.  =  0.64  of  the  total  cross-sectional  area, 
the  requisite  total  area  =  224.5  -5-  0.64  =  351  sq.  ft.  Hence,  the  length 
of  each  side  of  the  square  base  of  the  checker  work  =  -\/351  =  18.7  ft.,  or, 
approximately,  18  ft.  8.5  in.  The  requisite  height  for  the  checker  work, 
then,  =  6,250  -j-  351  =  17.8  ft.,  or,  approximately,  17  ft.  9.5  in. 


SEC.  424]  METHODS  OF  RECOOLING  CONDENSING  WATER  361 


•/Nc 


IOTE. — THE  TOTAL  HEIGHT  OF  A  COOLING-TOWER,  of  the  type  speci- 
fied above,  would  be  given  by  the  sum  of  the  fan-height  +  the  height  of  the 
checker  work  -f  2ft.  for  the  height  of  a  distributing  trough  (Fig.  331)  +  about 
4  ft.  for  the  depth  of  the  sump  or  well.  If  the  tower  is  erected  at  the 
grourid  level,  the  sump  may  be  sunk  below  the  surface  of  the  ground. 
^/NOTE. — THE  HEIGHT  OF  THE  FAN-BLOWER  REQUIRED  FOR  A  COOL- 
ING-TOWER may  be  obtained  from  manufacturers'  tables  of  the  dimensions 
and  capacities  of  such  blowers.  Typical  related  data  pertaining  to  fan- 
draft  towers,  for  use  in  connection  with  condensing-engine  plants,  are 
given  in  Table  424. 

424.  Table  Of  Related  Data  Pertaining  To  Forced-Draft 
Cooling-Towers  For  Use  With  Condensers  Of  Compound 
Condensing  Engines. 


Capacity 
of  con- 
denser, in 

Height 
of  cool- 

Dimensions 
of  cooling- 

Number 
and  size 

Speed  of 
Fans  in 
Revolu- 

Power re- 
quired 
for  Fan 

horse 
power 

ing  tower, 
in  feet 

tower  at 
base,  in  feet 

in  feet, 
of  fans 

tions  per 
min. 

in  horse 
power 

50 

25 

19      X  19.5 

1  -  6 

110 

1.25 

75 

25 

19.8  X  20.0 

1  -  6 

160 

1.75 

100 

25 

20.0  X  20.8 

1  -  7 

145 

2.25 

150 

25 

21.5  X  22.5 

1-8 

145 

3.50 

200 

25 

23.3  X  24.5 

1  -  9 

135 

5.50 

250 

26 

24.5  X  25.3 

1.-  10  ' 

135 

8.00 

300 

26 

26.5  X  27.0 

1-10 

145 

11.00 

400 

27.5 

27.5  X  24.5 

1  -  12 

115 

14.00 

500 

27.5 

29  X  30 

1  -  12 

145 

18.00 

425.  The  Cost  Of  A  Cooling-Tower,  erected  in  place,  may 
(PRACTICAL  ENGINEER,  1916)  be  from  $6  to  $7  per  kilowatt 
of  the  power  developed  by  the  plant.  Or,  otherwise,  from 
$4.50  to  $5.50  per  horse  power  of  the  engines  to  be  served. 
These  values  are  based  on  the  assumption  of  a  26-in.  vacuum 
in  condenser  operation. 

QUESTIONS  ON  DIVISION  10 

1.  Why  is  recooling  of  condensing- water  desirable? 

2.  What  phenomena  are  employed  in  the  recooling  of  condensing-water? 

3.  What  factors  determine  the  effectiveness  of  recooling  apparatus? 

4.  What  is  relative  humidity? 

5.  How  is  the  relative  humidity  of  the  air  determined  in  practice? 


362  STEAM  POWER  PLANT  AUXILIARIES         [Div.  10 

6.  Which  is  most  conducive  to  the  recooling  of  condensing-water — high  or  low  relative- 
humidity?     Why? 

7.  What  three  devices  or  methods  are  commonly  used  for  re-cooling  condensing-water? 
Under  what  conditions  would  each  be  most  advantageous? 

8.  Explain  the  operation  of  a  spray-fountain. 

9.  What  is  the  per  cent,  of  water-loss  from  a  spray-fountain,  relative  to  the  amount 
of  recooling  effected? 

10.  How  may  spray-fountains  be  protected  from  water-loss  by  high  winds? 

11.  What  is  the  usual  depth  of  cooling-ponds? 

12.  Can  spray-fountains  be  used  where  ground  space  is  unavailable?     How?    Explain. 

13.  What  per  cent,  of  the  total  power  developed  by  the  plant  is  required  for  spray- 
fountain  operation? 

14.  How  may  the  power  required  for  elevating  the  condensing-water  to  an  overhead 
spray-fountain  be  compensated  for? 

15.  What  are  the  essential  principles  of  cooling-tower  operation? 

16.  What  are  the  four  general  classes  of  cooling-towers? 

17.  What  average  per  cent,  of  efficiency  may  be  obtained  in  cooling-tower  operation? 

18.  How  does  the  water-loss  from  a  cooling-tower  compare  with  that  from  a  spray- 
fountain? 

19.  What  per  cent,  of  the  re-cooling  in  a  cooling-tower  is  generally  due  to  evaporation? 
How  is  the  remaining  per  cent,  of  the  re-cooling  effected? 

20.  In  what  respect  does  an  atmospheric  cooling-tower  differ  from  a  natural-draft 
closed  cooling-tower? 

21.  What  advantages  result  from  arranging  a  cooling-tower  so  that  it  may  be  used 
with  either  forced  or  natural  draft? 

PROBLEMS  ON  DIVISION  10 

vl.  The  air  entering  a  cooling-tower  has  a  dry-bulb  temperature  of  70  deg.  fahr. 
and  a  wet-bulb  temperature  of  60  deg  fahr.  The  air  leaving  the  tower  has  a  dry-bulb 
temperature  of  90  deg.  fahr.  and  a  wet-bulb  temperature  of  88  deg.  fahr.  What  is 
the  relative  humidity  in  each  case?  What  weight  of  water  does  the  air  absorb,  per  cubic 
foot,  while  passing  through  the  tower? 

2.  The  quantity  of  water  circulated  through  the  steam  condensers  of  a  1,000  h.p. 
engine  plant  is  40  Ib.  for  each  pound  of  steam  condensed.     The  engines  consume  15 
Ib.  of  steam  per  horse  power  per  hour.     What  should  be  the  area  of  a  simple  cooling 
pond  to  re-cool  the  condensing  water  in  summer?     What  should  be  the  area  if  the  pond 
were  equipped  with  a  spray-fountain? 

3.  In  Problem  1  the  air  re-cools  800  gal.  of  condensing  water  per  minute  through  20 
deg.  fahr.     The  water  enters  the  tower  at  105  deg.  fahr.  and  leaves  at  85  deg.  fahr. 
It  is  assumed  that  20  per  cent,  of  the  heat  abstraction  is  due  to  convection  and  radia- 
tion, while  the  remaining  80  per  cent,  is  due  to  evaporation.     What  volume  of  air  flows, 
per  minute,  through  the  tower?     What  is  the  efficiency  of  the  tower?     What  is  the  per 
cent,  of  evaporation-loss? 

^•i.  Assuming  that  the  cooling-tower  of  Problems  1  and  3  is  furnished  with  a  cypress- 
board  checker  work  (Fig.  329),  what  is  the  free  area  through  the  tower?  If  the  checker 
work  is  of  square  cross-section,  what  are  its  base-dimensions? 

5.  A  spray-fountain,  fitted  with  2-in.  nozzles,  is  to  operate  under  a  pressure  of  6  Ib. 
per  sq.  in.  The  quantity  of  water  circulating  through  the  condensers  is  40,000,000 
gal.  per  day  of  24  hr.  How  many  nozzles  are  needed?  What  pond  area  is  required? 


DIVISION  11 


STEAM-PIPING  OF  POWER  PLANTS 

426.  The  Steam-Piping  Of  A  Power  Plant  Generally  Com- 
prises   Two    Separate   Systems:  (1)  The    live-steam   piping. 
(2)   The    exhaust-steam    piping.     The    live-steam    piping    is 
usually  designed  to  convey  live  steam,  either  saturated  or 
superheated,  from  boilers  to  engines  and  other  steam-con- 
suming apparatus  at  pressures  from  about  100  to  300  Ib.  per 
sq.  in.     It  is,  therefore,  built  of  the  heavier  and  stronger  grades 
of   pipe    and  fittings.     The  exhaust-steam  piping  is  usually 
designed  to  carry  exhaust-steam,  from  turbines,  reciprocating 
engines  and  from  steam  pumps,  under  pressures  ranging  from 
less  than  atmospheric  to  perhaps  10  Ib.  per  sq.  in.     It  may, 
therefore,  be  built  of  comparatively  light  pipe  and  fittings. 

427.  The    Materials   For  Steam -Piping  comprise  mainly: 

(1)  Wrought  iron.     (2)  Mild  steel.     (3)  Cast-steel.     (4)  Cast- 
iron.     (5)  Malleable  iron.     Wrought-iron  pipe  is  much  favored 
on  account  of  its  reputation  for  ductility  and   durability. 
Pipe  made  of  mild  steel  produced  by  the  open-hearth  process 
is,  however,  commonly  con- 
ceded   to   be  equal  in  all 

respects  to  wrought-iron 
pipe.  Cast-steel,  cast-iron 
and  malleable  iron  are  used 
mainly  in  the  making  of 
fittings. 

428.  The    Grades    Of 
Steel    And   Wrought-iron 
Pipe     are:    (1)     Standard. 

(2)  Extra   heavy.     (3)  Double  extra  heavy    (Fig.    338).     The 
thickness,  and  weight  per  unit  of  length,  of  the  three  grades 
of  pipe  increase  in  somewhat  irregular  ratios. 

EXAMPLE. — The  thickness  of  a  5-in.  pipe  (Fig.  338)  advances  from 
0.247-in.  in  the  standard  grade  to  0.355-in.  in  the  extra  heavy  grade, 
and  0.71-in.  in  the  double  extra  heavy  grade.  The  weight  of  a  5-in. 

363 


•556-in.  ----H     K---S.5S-in. 


-H    K--5.56-m.--M 
1    I  I 


I-Stomotaroi         I-Extra  Heavy 


FIG.  338. — Inside  And  Outside  Diameters  Of 
Three  Grades  Of  Wrought  Iron  5-In.  Pipe. 


364 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


pipe,  per  foot  of  length,  advances  from  about  12.5  Ib.  in  the  standard 
grade  to  17.6  Ib.  in  the  extra  heavy  grade,  and  32.5  Ib.  in  the  double  extra 
heavy  grade.  Approximately  similar  ratios  are  noted  throughout  the 
tables  of  sizes. 

NOTE. — THE  SIZES  OF  ALL  STEEL  AND  WROUGHT  PIPES,  up  to  12-in. 
refer  to  the  nominal  inside  diameters.  Above  12-in.,  the  sizes  refer  to 
the  actual  outside  diameters.  These  large  pipes  are  made  in  several 
thicknesses  from  Y±  in.  to  1  in.  The  thinner  pipes  are  used  for  the  lower 
pressures  and  the  thicker  for  the  higher  pressures.  In  purchasing,  this 
large  pipe  is  specified  by  both  its  outside  diameter  and  thickness. 

429.  The  Grades  Of  Pipe  Fittings  Commonly  Used  In 
Steam -Piping  Systems  are:  (1)  Standard  cast-iron.  (2) 

B-3  in.  Open  Petwrn  C-Medium  -   D-Eccentrc 
A-Y-Benol;   Bewk    Sweep  Double     Tee--, 
-^-xjrcmch  Elbows.-  n 

e 


1  Inclic<^tir>of 
Stanc/artf  Fitting 


rtale  End 


FIG.  339. — Standard  Cast-Iron  Fittings. 


FIG.    340.  —  Standard    Malleable 
Fittings. 


Iron 


Standard  malleable  iron.  (3)  Extra  heavy  cast-iron.  (4) 
Extra  heavy  malleable  iron.  (5)  Extra  heavy  cast-steel.  (6)  Low- 
pressure  cast-iron.  Standard  cast-iron  fittings  (Fig.  339) 
are  designed  for  steam  pressures  up  to  125  Ib.  per  sq.  in. 


.-I-B.se 


.'X**** 


.       **          • 
Hngj  5weep:-''  Kurt'     Taper  Reducer- 


Fia.  341.—  Extra  Heavy  Cast-iron 
Fittings. 


—  -Collar 


FIG.  342. — Extra  Heavy  Malleable  Iron 
Fittings. 


Standard  malleable  iron  fittings  .(Fig.  340)  may  be  used 
for  steam  pressures  up  to  150  Ib.  per  sq.  in.  Extra  heavy 
cast-iron  fittings  (Fig.  341)  are  intended  to  withstand  steam- 
pressures  up  to  250  Ib.  per  sq.  in.  Extra  heavy  malleable 


SEC.  430]          STEAM-PIPING  OF  POWER  PLANTS 


365 


iron  fittings  (Fig.  342)  are  safe  for  steam  pressures  up  to 
250  Ib.  per  sq.  in.  Extra  heavy  cast-steel  fittings  (Fig.  343) 
are  safe  under  a  steam-pressure  of  350  Ib.  per  sq.  in.  and  a 
total  steam-temperature  of  800  deg.  fahr.  Thus  they  are 
available  for  use  in  piping  for  superheated  steam.  Low- 


>-l  -  Reo/wc  ing  Double 
Sweep  Tee 


•rl -Reducing 
'v,  Cross  Elbow-. 


5^3    «"*•') 

Raised  Face-' 


FIG.  343.  —  Extra    Heavy 
Cast-Steel  Fittings. 


•Straight  Face 


Fia.  344.  —  Low-Pressure  Cast- 
iron  Fittings. 


pressure  cast-iron  fittings  (Fig.  344)  are  suitable  for  steam 
pressures  up  to  25  Ib.  per  sq.  in.  They  may  be  used  in 
exhaust-steam  systems.  Their  use  in  live-steam  systems, 
even  where  the  pressure  does  not  exceed  25  Ib.  per  sq  in.,  is 
inadvisable. 

430.  The  Pipes  Commonly  Used  In  Steam-Piping  Systems 
Are  Classified  According  To  Three  Different  Types  Of  Con- 


Circular  Die-> 


•hecf 


M 


Skelp.' 


Weld- 


Fia.  345.— Method  Of  Forming  Butt-Welded  Pipe. 

struction:  (1)  Butt-welded.  (2)  Lap-welded.  (3)  Riveted. 
In  the  making  of  butt-welded  pipe  (A-Fig.  345)  the  squared 
edges  of  the  skelp,  B,  are  brought  to  a  welding  heat.  The 
end  of  the  pipe  is  then  formed  C-  and  the  edges  are  pressed 
together  D-  by  drawing  the  skelp  through  a  circular  die; 


366 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


M.  In  the  making  of  lap-welded  pipe,  the  edges  of  the  skelp 
are  scarfed  (A -Fig.  346)  and  the  skelp  is  rolled  into  tubular 
form.  The  skelp  is  then  brought  to  a  welding  heat  and  is 


FIG.  346.— Method  Of  Forming  Lap-Welded  Pipe. 

passed  (B-Fig.  346)  through  a  circular  groove  in  the  welding 
rolls.  The  weld  is  made  by  squeezing  the  overlapped 
scarfed  edges  together  between  the  walls  and  a  cast-iron 


xst  Iron.Fvraed  breel.or 
ffer/feosbfe  Inn  Riveted  Flanges  - 


FIG.  347.  —  Straight-Riveted  Steel  Pipe. 


mandrel.  Riveted  pipe  (Figs.  347  and  348)  is  made  of  sheet 
steel.  It  may  be  used  for  exhaust-steam  mains.  It  should 
not  be  used  in  live-steam  systems. 


-Forged!  Steel  Riveteol  Flanges 


TTITTTTYT 

"^"••--SpTroil  Seams  -- 

FIG.  348.— Spiral-Riveted  Steel  Pipe. 

NOTES. — THE  STRENGTH  OF  A  BUTT- WELD  is  about  73  per  cent,  of 
the  strength  of  the  plate  which  it  joins.  The  ultimate  strength  of  a 
butt-weld  in  a  steel  pipe  is  about  41,000  Ib.  per  sq.  in.  The  ultimate 
strength  of  a  butt-weld  in  a  wrought-iron  pipe  is  about  29,000  Ib.  per 
sq.  in. 


SEC.  431]          STEAM-PIPING  OF  POWER  PLANTS  367 

THE  STRENGTH  OF  A  LAP- WELD  is  about  92  per  cent,  of  that  of  the 
plate  which  it  joins.  The  ultimate  strength  of  a  lap-weld  in  a  steel  pipe 
is  about  52,000  Ib.  per  sq.  in.  The  ultimate  strength  of  a  lap  weld  in  a 
wrought-iron  pipe  is  about  31,000  Ib.  per  sq.  in.  Lap-weld  pipe  may  be 
used  for  all  purposes  of  live-steam  piping.  It  is  from  40  to  45  per  cent, 
more  expensive  than  butt-weld  pipe. 

431.  The  Trade  Meanings  Of  "Wrought-iron  Pipe"  And 
"Steel  Pipe"   are   not   generally  understood.     Steel   pipe  is 
(POWER,   Dec.    14,    1920,   page  948)   commonly  known    and 
billed  in  the  trade  as  "  wrought  pipe."    Jobbers  and  contract- 
ors are  prone  to  install  steel  pipe  instead  of  the  more  expensive 
wrought-iron  material  even  when  the  latter  is  specified.     They 
are  able  to  make  the  case  in  court  on  the  plea  "  wrought-iron 
pipe"  is  a  trade  term  meaning  either  wrought-iron  or  steel 
pipe  as  distinguished  from  cast-iron  pipe.     The  Executive 
Committee  and  Advisory  Board  of  the  National  Pipe  and 
Supplies  Association,  in  order  to  prevent  the  confusion  which 
is  heretofore  existed,  recommends  the  terminology  employed 
by  the  American  Society  for  Testing  Materials:  (1)  Welded 
wrought-iron  pipe.     (2)  Welded  steel  pipe.     If  this  standard 
terminology  is  followed  the  meanings  then  are: — (1)   That 
welded  pipe  is  pipe  which  is  welded  no  matter  what  it  is  made 
of.     (2)  That  welded  steel  pipe  is  pipe  made  by  welding  steel. 
(3)  That  welded  wrought-iron  pipe  is  pipe  that  is  made  of 
wrought  iron  by  the  welding  process.     (4)  That  wrought-iron 
pipe  is  pipe  made  of  wrought  iron  regardless  of  the  process  of 
manufacture. 

432.  The  Safe  Working  Pressures  For  Standard  Wrought 
Iron  And  Steel  Pipe  from  data  by  CRANE  Co.  are  as  follows: 
J^j  in.  to  J^  in.  butt  welded,  900  Ib.  per  sq.  in. ;  %  in.  to  1  in. 
butt  welded,  750  Ib.  per  sq.  in.;  1  in.  to  3  in.  butt  welded,  400 
Ib.  per  sq.  in.;  3J^  in.  to  5  in.  lap  welded,  400  Ib.  per  sq.  in.; 
6  in.  to  12  in.  lap  welded,  250  Ib.  per  sq.  in. 

NOTE. — More  conservative  practice  is  to  limit  steam  pressures  on  all 
standard  weight  pipe  to  250  Ib.  per  sq.  in.  Lap-welded  pipe  is  considered 
somewhat  more  reliable  than  butt-welded  and  is,  in  general,  preferred  for 
all  steam  piping  regardless  of  the  pressure.  Some  engineers  specify  only 
lap-welded  pipe  for  all  steam-power-plant  work. 


368 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


433.  Table  Showing  Good  Practice  Regarding  Grades  Of 
Pipe  For  Steam-Power  Plant  Installations.  All  pipe  for  pres- 
sures over  125  Ib.  per  sq.  in.  should  be  lap-welded.  (Con- 
densed from  Crane  Co.  specifications.) 


Pressure, 
Ib.  per 
sq.  in. 
gage 

Service 

Pipe  size, 
in  inches 

Grade  of  pipe 

Material 

Up  to  125 

Saturated  steam. 

Up  to  12  in  

Standard  Merchant  wt. 

14  to  18  in 

21  e  in   thick 

Steel 

Over  18  in  

H  in.  thick. 

125-200 

Saturated  steam 

Up  to  12  in 

Over  12  in  

%  in.  thick. 

200-250 

Saturated  steam 

Up  to  12  in 

Ox         1 

Over  12  in  

Ke  or  H  in.  thick. 

Steam  superheated 

Up  to  8  in  

Extra-strong. 

Steel 

Over  8  in  

J-^  in.  thick. 

Exhaust  steam. 

Up  to  12  in  

Standard  Merchant  wt. 

Steel 

14  to  24  in  

At  least  Me  in.  thick. 

Screwed 


.-Companion  flanges  i 


NOTE. — "STANDARD"  PIPE  (Sec.  428)  Is  MANUFACTURE?/  IN  Two  WEIGHTS:  (1) 
Full  card  weight.  (2)  Merchant  weight.  Full-card-weight  pipe  is  manufactured  to  con- 
form exactly  to  the  standard  dimensions.  Merchant-weight  pipe  is  not,  strictly,  quite 
as  thick  and  strong  as  is  full-card-weight  pipe. 

434.  Two  Principal  Types  Of  Joints  Are  Commonly  Used  In 
Steam  -Piping :  (1)  Screwed  joints.  (2)  Flanged  joints.  Screwed 
joints  (A -Fig.  349)  between  pipe-ends  and  fittings  are  usually 
recommended  for  steam-piping  where  the  pipe-size  does  not 
exceed  2.5-in.  This  however  depends  largely  on  the  pressure 
and  service  for  which  the  pipe  is  to  be  used;  generally,  for 

pressures  below  125  Ib.  per  sq. 
in.  the  piping  connections  are 
"  sere  wed "  only  for  pipes  up  to 
2^2  in-  nominal  diameter. 
Flanged  joints  (£-Fig.  349)  are 
FIG.  349.— screwed  And  Flanged  Joints  generally  easier  to  manipulate 

In  Steam-Piping.  *  .    .  r 

than  are.  screwed  joints.     They 

afford  ready  means  for  disconnecting  the  various  sections  of 
a  piping-system.  Their  use  is  recommended  in  all  steam- 
piping  larger  than  2.5-in. 

NOTE. — Flanges  commonly  form  screwed  joints  with  the  pipe-ends. 
Hence,  the  construction  of  a  flanged  joint  in  a  pipe-line  may  and  usually 
does  entail  (B-Fig.  349)  the  use  of  one  subsidiary  screwed  joint. 


Screwed 


{Joints- 


SEC.  435] 


STEAM-PIPING  OF  POWER  PLANTS 


369 


,-Thread         ..--Chamfer 


435.  The  Principal  Methods  of  Attaching  Flanges  To  Pipe- 
Ends  are  :  ( 1 )  Threading.  ( 2 )  Shrinking.  ( 3 )  Flaring  or 
lapping.  (4)  Welding.  Threading  (/,  Fig.  350)  consists  in 
screwing  the  flange  on  the  pipe-end.  Strength  and  lightness 
are  insured  by  forcing  on  the  flange  until  the  pipe-end  projects 
beyond  the  flange-face.  The  pipe-end  is  then  cut  off  flush  with 
the  flange-face.  In  the  shrinking  method  (II,  Fig.  350)  the 
pipe-end  is  turned  truly  cylindrical.  The  flange  is  bored  to  a 
shrink-fit  and  the  face-end  of  the  bore  is  chamfered.  The 
flange  is  then  heated  to  redness,  and  is  slipped  over  the  pipe- 
end  until  the  end  projects  beyond  the  flange-face.  When  the 
flange  has  cooled  somewhat,  the  pipe-end  is  beaded  into  the 
chamfer  with  a  ball-peen  hammer.  The  pipe-end  is  finally 
turned  off  flush  with  the  flange-face. 

In  the  flaring  or  lapping 
method  (III,  Fig.  350)  the 
flange  is  bored  slightly 
larger  than  the  outside 
diameter  of  the  pipe.  The 
end  of  the  pipe  is  flared 
or  belled.  An  abruptly 
flared  end  (III,  Fig.  350) 
is  called  a  lapped  end. 
The  flange  fits  loosely 
around  the  pipe  and  forms 
a  swivel-joint  with  the  lap.  m-nored or  Lapped 

This  imparts  flexibility  tO  FIG-  350. — Methods  Of  Securing  Companion 
.1  i  ,  i  Flanges  To  Pipe  Ends. 

the     structure    when    the 

flange  is  bolted  tightly  to  a  flanged  fitting.  One  method  of 
welding  (IV,  Fig.  350)  consists  in  heating  both  the  flange  and 
pipe-end  to  a  welding  heat  and  squeezing  them  together 
under  heavy  pressure,  into  a  single  mass.  Flanges  may  also 
be  arc-welded  or  acetylene  welded  to  the  pipe-ends. 

NOTE. — Pipe-end  flanges  are  commonly  called  companion  flanges. 

NOTE. — The  cost  of  an  extra-heavy  forged-steel  welded  flange  being  re- 
garded as  a  basis  of  comparison,  or  as  100  per  cent.,  the  relative  costs 
of  other  types  of  attachment  of  extra-heavy  flanges,  made  of  different 
materials,  may  be  expressed  as  follows: 

24 


370 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


Forged-steel  shrunk,    120  per  cent.     Cast-steel  shrunk,  105  per  cent. 

Cast-iron  shrunk,  60  per  cent.     Forged  steel  flared  or  lapped,  95  per  cent. 

Cast-steel  flared  or  lapped,  75  per  cent.  Malleable  iron  flared  or  lapped, 

55  per  cent.  Cast-iron  flared 
or  lapped,  50  per  cent.  Forged- 
steel  threaded,  95  per  cent. 
Cast-steel  threaded,  60  per 
cent.  Malleable-iron  threaded, 
45  per  cent.  Cast-iron 
threaded,  25  per  cent. 

436.  Low-Resistance 
To  Steam-Flow  In  The 
Turns  Of  A  Piping  System 
Is  Facilitated  By  The  Use 
Of  Pipe-Bends  (Fig.  351). 
Bends  (Fig.  352)  are  also 
used  to  absorb  the  con- 
traction and  expansion 
movements  of  piping.  The  radii  of  pipe-bends  (R-Fig.  352) 
should  always  be  as  great  as  circumstances  will  permit.  The 


FIG. 


351.-Standard  Pipe-Bends  For  Making 

Tums  in  Piping  Systems. 


Fid.  352. — Bends  For  Taking  Up  Expansion  Stresses  In  Piping  Systems. 

longer  the  radius,  the  greater  the  flexibility  at  the  bend.  Also 
the  larger  the  bend  the  less  the  liability  of  buckling  the  pipe 
when  forming  the  bend. 


SEC.  437]          STEAM-PIPING  OF  POWER  PLANTS 


371 


NOTES. — THE  MINIMUM  ADVISABLE  RADIUS  FOR  A  PIPE-BEND  in  a 
steam-line  (.R-Fig.  352),  for  pipe  sizes  from  2.5-in.  to  16-in.,  is  five  times 
the  nominal  diameter  of  the  pipe. 

THE  MINIMUM  ADVISABLE  LENGTHS  OF  THE  TANGENTS  OR  STRAIGHT 
PARTS  OF  PIPE-BENDS  (Fig.  352)  when  the  companion  flanges  are  either 
threaded  (7  Fig.  350)  or  shrunk  (77,  Fig.  350)  on,  increases,  in  regular 
progression,  from  4-in.  for  2.5-in.  pipe  to  11-in.  for  9-in.  pipe  and  to 
18-in.  for  16-in.  pipe. 

When  the  flanges  are  flared  or  lapped  (777,  Fig.  350)  the  tangent- 
lengths  range  from  6-in.  for  2.5-in.  pipe  to  9-in.  for  9-in.  pipe  and  to 
18-in.  for  16-in.  pipe.  When  the  flanges  are  welded,  the  range  of  tangent- 
length  is  from  5-in.  for  2.5-in.  pipe  to  6-in.  for  9-in.  pipe  and  to  8-in.  for 
16-in.  pipe. 

437.  Three  Methods  Are  Available  For  Distributing  The 
Steam  Supplied  By  A  Boiler  Plant  which  consists  of  more  than 


'•-Stop  and  Check- 
'•Voices ' 


FIG.   353. — Single  Header  System  Of  Steam  Piping. 

one  boiler  unit:  (1)  The  single  header  (Fig.  353).  (2)  The 
loop  header  or  duplicate  headers  (Figs.  354  and  355).  (3)  The 
unit  group  (Fig.  356).  The  single  header  is  the  least  expensive 
to  install.  It  is,  however,  the  least  convenient  arrangement. 
Thus,  if  it  were  necessary  to  repair  the  section  of  main  piping 
between  boilers  C  and  D  (Fig.  353)  boilers  A,  B  and  C  would 
not  be  available  for  supplying  the  prime  movers  to  the  right 
of  the  defective  section,  nor  would  boilers  7),  E,  F,  G  and  H  be 


372 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


available  for  supplying  the  apparatus  to  the  left.  With  the 
duplicate  headers  (Figs.  354  and  355)  the  cross-connections 
and  the  arrangement  of  stop  valves  insures  unrestricted  use 
of  all  the  boilers  and  prime  movers,  even  though  it  be  neces- 
sary to  isolate  a  section  of  the  piping  for  repairs. 


Stop  emu  Check  Voilves 

Fio.  354. — Proper  Method  Of  Connecting  A  Set  Of  Boilers  To  Duplicate  Main  Headers 
Continuation  Through  Engine-Room  Of  Headers  X  and  F'ls  Shown  In  Fig.  355. 

.-Double- Sweep  Tees                      Stop-Valves  in  ..-Main  Headers 
...' -Cross- Connections- .,_-;_" 


>-'-,. -Quarter  Bends 


-..     ,-•        * 
!  Sfngte-Smep  Teet-^Jag*  : 


Steam  Pipe  W  .-Steam  Pipes   M  Blind 
fi$g>/f**d  ^.-toCompound  ^  to  Turbines-  ^T/anges 


Steam  Pipe 


--S/mpii  Engine 


Fio.   355. — Continuation  Of  Duplicate  Headers  X  and  Y,  Fig.  354,  Through  Engine 

Room. 

438.  With  The  Unit-Group  Arrangement  (Fig.  356)  each 
engine  is  piped  directly  to  an  individual  set  of  boilers,  usually 
four,  as  boilers  A,  B,  C  and  D,  or  E,  F,  G  and  #,  or  /,  J,  K 
and  L.  Equalizer  pipes  are,  however,  employed  to  bond  the 
piping  of  all  the  boilers  in  a  single  system.  These  equalizers 
are,  usually  of  the  same  size  as  the  main  pipes  leading  to  the 
engines. 


NOTE. — Pressure-equalization  throughout  the  system  is  the  sole  func- 
tion of  the  equalizer  or  header-pipes  (Fig.  356).     They  are  not  designed 


SEC.  439]          STEAM-PIPING  OF  POWER  PLANTS 


373 


to  provide  storage  space,  as  the  headers  in  the  older  arrangements 
(Figs.  353  and  354)  are,  in  a  measure,  required  to  do.  Hence,  it  is  par- 
ticularly advisable,  where  the  unit-group  method  of  piping  is  used, 
that  ample  receiver-separators  be  installed  close  to  the  engine  throttle- 
valves. 


;  Automate  stop---. 
vwntf  Check  Vy]ves\ 


FIG.  356. — Unit  System  Of  Main  Steam  Piping  Showing  Three  Unit  Groups. 

439.  Steam-Pipe  Sizes  May  Be  Determined  Graphically 

by    means    of    a    chart    (Fig.    357)    which    was    devised  by 
H.  V.  Carpenter. 

EXAMPLE. — Find,  graphically,  the  pipe-size  required  to  supply  30,000 
Ib.  of  steam  per  hour  to  an  engine  if  the  allowable  pressure  drop  be- 
tween engine  and  boiler  is  3  Ib.  per  sq.  in.  The  boiler  supplies  the 
engine  through  a  pipe  which  is  150  ft.  long.  The  operating  steam- 
pressure  is  185  Ib.  per  sq.  in.  gage. 


374 


STEAM  POWER  PLANT  AUXILIARIES 


SOLUTION. — The  given  rated  steam-flow  of  30,000  Ib.  per  hr.  may, 
allowing  a  50  per  cent,  excess  rating,  be  reduced  to  (30,000  -5-  60)  X 
1.50  =  750  Ib.  per  min.  The  given  gage  pressure,  185  Ib.,  is  equivalent 
to  (185  +  15  =)  200  Ib.  absolute  pressure.  Also,  the  given  pressure- 
drop  of  3  Ib.  in  150  ft.  corresponds  to  2  Ib.  in  100  ft.  From  A,  corres- 


ponding  to  750  Ib.  on  the  base  line  (Fig.  357)  proceed  vertically  upward 
to  B  on  the  line  of  200  Ib.  absolute  pressure.  Proceed  thence  downward, 
parallel  to  the  oblique  lines,  to  C  on  the  line  of  2  Ib.  pressure-drop. 
Tracing  vertically  upward  from  C,  the  point  of  intersection,  D,  with  the 
top  line  indicates  the  required  pipe  size  to  be  about  6.7  in.,  or  practically, 
7  in. 


SEC.  440]          STEAM-PIPING  OF  POWER  PLANTS  375 

440.  A  Simple  Formula,  For  Computing  The  Pipe   Size 
Necessary  To  Deliver  Steam  At  A  Given  Rate,  which  is  used 
often  in  practice  is  given  below.     In  using  this  formula,  a 
steam-flow  velocity,  which  practice  has  shown  will  not  induce 
an  excessive  pressure  drop,  is  assumed.     Then,  the  required 
pipe  diameter  or  area  may  be  obtained  by  substituting  the 
other  known  values : 

(97)  di  =  13.54A/yr-  (diam.  inches) 

(98)  Ai  =  —f. —  (area,  sq.  in.) 

L/l)m 

Wherein  di  =  actual  internal  diameter  of  pipe,  in  inches. 
W  =  equivalent  weight  of  steam  flowing  through  pipe,  in 
pounds  per  minute.  D  =  density  of  steam  at  the  given  pres- 
sure, in  pounds  per  cubic  foot.  vm  =  velocity  of  flow  of 
steam  in  pipe  (see  Sec.  441)  in  feet  per  minute.  At-  =  in- 
ternal area  of  pipe,  in  square  inches. 

NOTE. — THE  ABOVE  FORMULA  MAY  BE  USED  FOR  FIGURING  THE 
PIPE  SIZE  REQUIRED  FOR  A  RECIPROCATING  ENGINE  if  the  valve  cut  off 
is  known.  For  example,  if  30,000  Ib.  of  steam  is  used  by  the  engine 
per  hour  and  the  cut  off  is  ^,  then  the  equivalent  flow  will  be  appro- 
ximately: 4  X  30,000  =  120,000  Ib.  per  hr.  These  formulae  are  not  re- 
commended for  pipes  under  3  in.  in  diameter. 

EXAMPLE. — A  steam  engine  which  is  set  for  Y$  cut  off  requires  12,000 
Ib.  of  steam  per  hr.  The  steam  pressure  is  125  Ib.  per  sq.  in.  gage.  A 
velocity  of  6,500  ft.  per  min.  is  allowable  in  the  pipe.  What  size  of  pipe 
is  required?  SOLUTION. — The  steam  velocity  is  equivalent  to  that  when 
the  steam  flows  continuously  at  the  rate  of  3  X  12,000/60  =*  600  Ib.  per 
min.  Substituting  in  the  above  formula :  di  =  13.54\/W/Z)vTO  =  13.54  X 
V(  600  )  -5-  (0.3107  X  6,500)  =  7.3  in.  internal  diameter  or  a  7  in.  pipe 
is  sufficiently  large  if  not  too  long  (Sec.  444). 

NOTE. — This  size  pipe  will  have  a  maximum  pressure  drop  as  computed 
by  Fig.  357  of  4  Ib.  per  sq.  in.  per  100  ft. 

441.  The  Allowable  Steam-Flow  Velocities  Used  In  Practice 

are  about  as  follows:  For  average  power-plant  installations: 
saturated  steam,  6,000  to  8,000  ft.  per  min.  superheated 
steam,  8,000  to  12,000  ft.  per  min.  exhaust  steam  4,000  ft. 
per  min.  In  large  stations  the  velocity  may  be,  for  superheated 
steam,  14,000  ft.  per  min.  for  reciprocating  engines  and  15,000 
ft.  per  min.  for  turbines.  In  one  large  eastern  turbine  station 


376  STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 

the  velocity  is  21,000  ft.  per  min.  In  another  reciprocating- 
engine  installation  a  velocity  of  15,000  ft.  per  min.  is  used 
without  any  apparent  adverse  effect  on  economy. 

442.  The  Drops  In  Pressure  In  Steam  Mains  Allowed  In 
Practice  range  up  to  perhaps  30  Ib.  per  sq.  in.  from  boiler 
to  engine.     A  total  loss  in  pressure  of  more  than  15  per  cent., 
however,  is  not  recommended  although  the  friction  in  the 
mains  causes  heat  which  superheats  the  steam  and  does  not 
represent   actual   energy   lost.     Some   engineers   recommend 
less  than  4  Ib.  per  sq.  in.  drop  in  pressure  per  100  ft.  of  pipe. 
From  Sec.  445  it  will  be  noted  that  the  loss  in  pressure  due  to 
a  valve  is  large  compared  to  that  in  100  ft.  of  straight  pipe. 

NOTE. — Where  large  receiver-separators  are  installed  close  to  engine 
throttle  valves,  pressure-drops  of  from  1.5  to  2.5  Ib.  per  100  ft.  of  pipe  are 
permitted.  The  corresponding  velocity  of  steam-flow  is  about  9,000  ft. 
per  min.  Where  a  pipe-line  is  very  long,  the  pressure-drop  per  100  ft. 
must,  obviously,  be  kept  low  in  order  that  a  fair  percentage  of  the  initial 
steam-pressure  may  be  realized  at  the  place  of  delivery. 

443.  The  Average  Pressure-Drop  In  Exhaust-Steam  Main 
Piping  is,  ordinarily,  from  about  0.2  to  0.4  Ib.  per  100  ft.  of 
pipe  where  the  engines  are  run  non-condensing.     It  is  from 
about  0.2  to  0.4  in.  of  mercury  column  per  100  ft.  of  pipe 
where  the  engines  are  run  condensing  with  a  vacuum  of  about 
26  in. 

444.  The  Size  Of  A  Main  Pipe  Having  A  Carrying  Capacity 
Equal  To  The  Combined  Capacities  Of  Two  Or  More  Branch 
Pipes  May,  for  the  same  velocity  of  steam-flow  and  other 
conditions,  be  formed  by  the  following  formula : 


(99)  dim  =  Vdn2  +  di22  +  drf  +  etc.          (inches) 

Wherein  dim  =  actual  inside  diam.,  in  inches,  of  main  pipe; 
dti,  diz,  dis,  etc.  =  inside  diameters,  in  inches,  of  branch  pipes. 

EXAMPLE.  —  Assuming  the  same  velocity  of  steam-flow  in  the  main  and 
branch  piping,  what  should  be  the  size  of  a  header  to  supply  four  branches 
having  diameters  of  3-in.,  3^-in.,  5-in.,  and  6-in.,  respectively? 

SOLUTION.  —  By  For.  (99),  dim  = 


_ 

+  drf  +  drf  +  dt42  =  \/32  +  3.52  +  52  +  62  =  9.1-in. 
Hence,  a  10-in.  pipe  is  required,  since  this  is  the  next  larger  size  to  the 
value  found. 


SEC.  445]          STEAM-PIPING  OF  POWER  PLANTS 


377 


445.  The  Pressure  -Drop  Due  To  The  Presence  of  Globe 
Valves  And  Right-Angled  Fittings  In  Steam  Pipes  may  be 

taken  into  account,  in  the  computations  for  size,  by  applying 
Briggs  formulae,  which  are  as  follows  : 


(100) 
(101) 


Lv  = 


i  +  (l  + 
Le  =  7Qdi  -f-   (l  +  ^ 


(inches) 
(inches) 


(1  +  (3.6/d 


114  X  7  * 


Wherein  Lv  =  pipe-length,  in  inches  having  resistance  equiva- 
lent to  that  of  one  globe  valve,  Le  =  pipe-length,  in  inches, 
having  resistance  equivalent  to  that  of  one  standard  90  deg. 
elbow,  di  =  internal  diameter  of  pipe,  in  inches. 

NOTE.  —  GATE  VALUES  AND  PIPE  BENDS  PRODUCE  PRESSURE  DROPS 
only  equal  to  the  length  of  pipe  they  actually  contain.  That  is,  a  pipe 
bend  which  is  30  in.  long  measured  along  its  circular  center  line  would 
produce  the  same  pressure  drop  as  a  straight,  30-in.  length  of  the  same- 
size  pipe;  a  gate  valve  measuring  8  in.  from  face  to  face  would  introduce 
the  same  pressure  drop  as  an  8-in.  straight  length  of  pipe  of  the  size  into 
which  the  valve  is  designed  to  be  fitted.  These  drops  may  be  read  from 
Fig.  357. 

EXAMPLE.  —  To  how  many  feet  of  pipe-length  would  the  resistance 
offered  by  one  globe  stop-valve  and  two  standard  90-deg.  elbows  in  a 
7-in.  steam-line  be  equivalent? 

SOLUTION.—  By  For.  (100),  Lv  = 
[1  +  (3.6  -f-  7)]  in.  =  527.08  =  527.08 
-5-12=  43.9  /t.  By  For.  (101),  Le  = 
7Qd  -s-  (1  +  3.6M-)  =  76  X  7  +  (1  + 
3.6  -4-  7)  =  351.39  in.  =  351.39  -f-  12  = 
29.28  ft.  Hence,  the  total  equivalent 
pipe-length  =  43.9  +  29.28  X  2  = 
102.46  ft. 

446.  Linear  Expansion  In 
Steam  Pipes  tends  to  produce 
bending,  buckling,  and  tensile 

&'  . 

Stresses    m    the    piping.       Strains 

due   to    these    stresses  are   ob- 

viated  (Figs.  352,  358,  359  and  360)  by  the  use  of  compensating 

devices. 

NOTE.  —  EXPANSION  SLIP-JOINTS  (Fig.  359)  are  mainly  used  with  very 
large  pipes,  and  where  space  prohibits  (Figs.  352  and  358)  long-radius 


FIG.  358.  —  Double-Swing   Or  Swivel 
Joint  For  Taking  Up  Expansion  In  Pipe 

Lines- 


378 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


bends,  or  swivel  joints.  When  slip-joints  are  necessary,  binding  in  the 
joint,  due  to  sagging  of  the  pipe,  must  be  guarded  against  by  erecting 
substantial  supports  at  each  end.  Also,  the  pipe  must  be  securely  an- 
chored to  prevent  the  steam-pressure  from  forcing  the  joint  apart. 


.-'Joint      ,'Corruyotfeol  Copper 
'.     xTTTl :  Casing^  for  Taking  up 
,r shncf  Pipe  Line 


^.-Horizontal  ftun  of  Pipe--., 
(Packing  Glands*. 


Sleeve--'     x'  Anchor  Base- '\  'v5/eeve 
'Fibrous  Packing-' 

FIG.  359. — Double-Slip  Expansion  Joint. 


'Polished     ''Sfeel  Reinforcing 
Lining 


Rings 


FIG.  360. — Corrugated  Expansion  Joint. 


447.  The  Linear  Expansion  Occurring  In  Steel  And  Wrought- 
Iron  Steam  Pipes  may,  for  given  lengths  of  piping  and  ranges 
of  temperatures,  be  found  by  the  following  formula: 

(102)  I  =  eiLTf  (inches) 

Wherein:  Z=  the  linear  expansion  of  the  pipe,  in  inches,  ei  = 
the  coefficient  of  linear  expansion  (see  note  below).  L  =  the 
original  length  of  steam  pipe,  in  inches.  T/  =  the  change  of 
temperature,  in  degrees  fahrenheit. 

NOTE. — The  coefficient  of  linear  expansion  (et)  for  charcoal  iron  is 
0.000,006,86;  Bessemer  steel,  0.000,006,99;  seamless  open-hearth  steel, 
0.000,006,88;  cast  iron,  0.000,006,2;  cast  steel,  0.000,006. 

EXAMPLE. — What  will  be  the  linear  expansion  in  a  straight  150  ft. 
line  of  Bessemer  steel  pipe  when  steam  at  a  pressure  of  125  Ib.  per  sq. 
in.,  gage,  is  admitted,  if  the  pipe  has  a  temperature  of  60  deg.  fahr.  at 
the  time  of  erection?  SOLUTION. — A  table  (see  the  author's  PRACTICAL 
HEAT)  of  the  properties  of  saturated  steam  gives  the  temperature  at 
125  Ib.per  sq.  in.  gage  as  353.1  deg.  fahr.  By  For.  (102)  I  =  etLTf  = 
0.000,006,99  X  (150  X  12)  X  (353.1  -  60)  =  3.69  in. 

448.  The  Least  Length  Of  Pipe  Necessary  For  A  Bend  Or 
Loop  To  Take  Up  The  Expansion  In  A  Run  Of  Pipe  Of  Given 
Length  may  be  found  by  Rayne's  formula,  which  is  as  follows : 

(103)  Lb  =  OM3VdoLpTf  (feet) 
Wherein:  Lb  =  least  length,  in  feet,  of  pipe  required  for  bend. 
d0  =  external    diam.,    in    inches,    of   pipe.    Lp  =  length,    in 


SEC.  449]          STEAM-PIPING   OF  POWER  PLANTS 


379 


feet,  of  pipe-line, 
heit. 


Tf  =  temperature  rise  in  degrees  Fahren- 


EXAMPLE. — What  is  the  least  length  of  pipe  that  should  be  used  in 
making  a  double-offset  expansion  U-bend  (A,  Fig.  352)  to  be  installed  in 
a  straight  150-ft.  run  of  6-in.  pipe  designed  to  carry  steam  at  150  Ib. 
pressure,  gage,  if  the  temperature  of  the  piping  when  erected  is  60  deg. 
fahr.? 

SOLUTION. — A  table  of  the  properties  of  saturated  steam  gives  the  tem- 
perature at  150  Ib.  pressure,  gage,  or  165  Ib.  pressure,  absolute,  as  366 
deg.  fahr.  The  outside  diam.  of  a  6-in.  pipe  is  6.625  in.  By  For.  (103)  Lb 
=  0.043  \/doLpTf  =  0.043  X  A/6.625  X  150  X  (366  -  60)  =23.7  ft. 

NOTE. — The  results  obtained  with  the  preceding  formula  can  be  applied 
directly  only  with  steam  pipes  of  the  smaller  sizes.  With  the  larger 
sizes,  it  may  be  necessary  to  increase  the  computed  lengths  in  order  to 
conform  to  the  minimum  allowable  ratio  (Sec.  436)  of  pipe-diameter  to 
radius  of  curvature,  and  to  the  prescribed  tangent  lengths. 

449.  Vibration  In  Steam-Piping  is  generally  caused  by  a 
pulsating  steam-flow.  The  pulsations  may  be  due  to  the 


,-Sfearn  Main 


.-Cast-Iron 


I -Beam-. 


Spring-. 


'•3  in.  Extra-Heavy  Pipes  for  Absorbing  ./' 
Lateral  Vibration  of  Steam  Main-- ' ' 

Fid.  361.  —  Devices  To  Prevent 
Transmission  Of  Pipe- Vibration.  A, 
Floor-Support  For  Use  Where  Space 
Is  Ample.  B,  Double-Spring  Hanger 
For  Use  Where  Head-Room  Is  Limited. 


In  Concrete  Floor 


FIG.  362.  —  Devices  To  Prevent  Trans- 
mission Of  Pipe-  Vibration.  A,  Floor-Support 
For  Use  Where  Space  Is  Restricted.  B,  Simple 
Spring  Hanger  With  Safety  Device. 


alternate  opening  and  closing  of  the  admission  valves  of  reci- 
procating engines.  Transmission  of  the  vibration  to  the 
foundations  and  walls  of  buildings  may  be  prevented  (Figs. 
361  and  362)  by  special  supporting  devices. 


380 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


450.  Various  Devices  Are  Used  For  Staying  And  Supporting 
Steam-Piping  in  order  to  prevent  deflection  and  vibration. 
These  devices  mainly  comprise:  (1)  Plain  hangers  (Fig.  363). 


< — I -Beam 


j  -Binoling- 

11  * 


FIG.    363.  — An    Ordinary 
Pipe  Hanger. 


Bracket— 


Wall- 


FIG.  364.— Wall-Bracket, 
With  Binding  Rolls,  For 
Supporting  Steam  Main. 


FIG.  365. — Simple  Floor- 
Stand  For  Supporting 
Steam  Main. 


(2)  Wall-brackets  (Fig.  364).  (3)  Floor  stands  (Fig.  365). 
(4)  Anchors  (Fig.  366).  (5)  Counter-balancing  hangers  (Fig. 
367).  Plain  hangers  should  be  free  to  swing  (Fig.  363)  in  the 


,  - Guides  (Secured  to  Wall) 

f'"  .' -Counterweights - 
f  -Steel  Levers- .. 


FIG. 


Wall-... 


366.  —  An     Ordinary    Pipe- 
Anchorage. 


FIG.  367. — Method  Of  Suspending  And  Counter- 
balancing Expansion  Loops  In  Steam  Mains. 


direction  of  the  length  of  the  pipe.  Also,  they  should  also  be 
provided  with  a  means  for  height-adjustment.  Wall-brackets 
with  roll-binders  (Fig.  364)  allow  for  free  linear  expansion  of 
the  pipe,  but  prevent  lateral  movement.  Such  binders  should 


SEC.  451]          STEAM-PIPING  OF  POWER  PLANTS 


381 


be  used  in  supporting  the  ends  of  horizontally-placed  long- 
radius  bends.  An  anchor  (Fig.  366)  is  designed  to  hold  the 
pipe  immovable,  at  the  place  of  anchorage,  against  expansion 
stresses.  Counter-balancing  hangers  (Fig.  367)  are  designed 
to  sustain  the  weight  of  expansion-loops,  while  giving  free 
play  to  the  rise  and  fall  of  the  loops  under  alternate  expansion 
and  contraction. 

451.  The  Heat  Losses  From  Bare  And  Insulated  Steam 
Pipe  are  as  follows  (based  on  Marks'  Mechanical  Engineers' 
Handbook) : 

Insulation  No.  1  is  of  a  hard  fire-proof  variety  of  asbestos  of  relatively 
poor  insulating  value.  No.  2  is  sponge-felted  asbestos.  The  conductivity 
of  most  insulation  for  pipes  is  intermediate  between  these  two  sets  of 
values.  The  insulation  is  assumed  to  be  about  1  in.  thick 


Temperature  difference,  pipe  and 
air,  deg.  fahr. 

50 

100 

200 

300 

400 

500 

Loss  in  B.t.u.  per  hr.  per 
deg.     fahr.     temperature 
difference   per   sq.    ft.    of 
pipe  surface. 

Bare 
pipe 

1.95 

2.15 

2.665 

3.26 

4.035 

5.18 

Insul- 
ation 
No.  1 

0.63 

0.65 

0.715 

0.781 

0.856 

0.967 

Insul- 
ation 
No.  2 

0.34 

0.35 

0.369 

0.391 

0.414 

0.439 

452.  The  Condensation  Due  To  Loss  Of  Heat  From  Bare 
Steam  Pipes  may  be  found  by  the  following  formula: 

2.7A,(Tf8-Tfa) 


(104) 


Wc  = 


(Ib.  per  hr.) 


Wherein:  Wc  =  weight  of  condensation,  in  pounds  per  hour. 
A/  =  area  of  external  surface  of  pipe,  in  square  feet.  Tf<  = 
steam  temperature  at  given  pressure,  in  degrees  fahren- 
heit.  T fa  =  temperature  of  surrounding  air,  in  degrees 
Fahrenheit.  Hv  =  latent  heat  of  steam  at  given  pressure, 
in  British  thermal  units  per  pound. 


382 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 


EXAMPLE. — The  external-surface  area  of  4-in.  pipe  is  1.178  sq.  ft.  per 
ft.  of  length.  What  will  be  the  quantity  of  condensation  in  40  ft.  of 
bare  4-in.  pipe  carrying  steam  at  105  Ib.  pressure,  gage,  when  the  sur- 
rounding air-temperature  is  60  deg.  fahr.? 

SOLUTION. — A  table  of  the  properties  of  saturated  steam  (Author's 
PRACTICAL  HEAT)  gives  the  temperature  of  steam  at  the  given  pressure 
as  341  deg.  fahr.,  and  the  latent  heat  as  877.2  B.t.u.  By  For.  (104), 
Wc  =  2.7Af(T/,  -  Tfa)  -r  Hv  =  2.7  X  1.178  X  40  X  (341  -  60)  i 
877.2  =  40.75  Ib.  per  hr. 


•"•<:!"-  -  '."-.-  .-."•'  v  :-":N      •••  Q-     Drain  Valve-- 
" 

FIG.   368. — The  Holly  Steam-Loop  For  Draining  High-Pressure  Piping. 

453.  Excessive  Loss  Of  Heat  From  Steam  Pipes  May  Be 
Prevented  by  covering  the  pipes  with  heat-insulating  material. 
Incombustible  mineral  substances,  as  magnesia  and  asbestos, 
are  commonly  used  for  this  purpose.  All  steam-pipe  cover- 
ings should  be  at  least  1-in.  thick.  The  heat-loss,  with  a  good 


SEC.  454]          STEAM-PIPING  OF  POWER  PLANTS  383 

covering,  may  be  reduced  to  about  15  per  cent,  of  that  occur- 
ring with  bare  pipe,  or  even  less. 

454.  The  Condensation  In  High-Pressure  Steam-Piping 
May  Be  Returned  To  The  Boilers  With  A  Holly  Loop  (Fig. 
368).  The  condensation  gravitates  to  a  receiver,  A, wherein 
it  is  broken  into  a  spray  by  passing  through  a  perforated 
plate.  Connections  should  be  made  to  the  receiver  from  all 
parts  of  the  piping  system  wherein  water  might  become 
pocketed.  Due  to  the  discharge  of  steam  from  the  discharge- 
chamber  C,  through  the  vent-pipe,  P,  and  reducing-valve, 
into  the  feed-water  heater,  the  pressure  in  the  discharge- 
chamber  is  less  than  that  in  the  receiver.  Hence,  a  current 
of  water-spray,  mixed  with  steam-vapor,  ascends  through  the 
riser  R.  The  steam  and  water  separate  in  the  discharge- 
chamber.  The  water  gravitates  to  the  boilers  through  the 
drop-leg  D.  The  discharge-chamber  is  placed  at  an  elevation 
that  will  insure  a  sufficient  hydrostatic  head  to  overcome  the 
excess  of  boiler  steam-pressure  over  the  discharge-chamber 
steam-pressure.  Circulation  in  the  loop  is  started  by  opening 
valve  S.  When  steam  appears,  valve  S  is  closed  and  the 
reducing  valve  is  opened. 

QUESTIONS  ON  DIVISION  11 

1.  What  pressures  are  commonly  carried  in  live-steam  piping?     In  exhaust-steam 
piping? 

2.  What  are  the  ordinary  materials  of  steam-piping? 

3.  Enumerate  the  regular  grades  of  steel  and  wrought-iron  pipe. 

4.  To  what  dimensions  do  the  nominal  sizes  of  piping  refer? 

5.  Enumerate  the  grades  of  pipe  fittings  commonly  used. 

6.  What  is  the  maximum  advisable  pressure  for  malleable-iron  fittings?     For  standard 
cast-iron   fittings?     For  extra    heavy   cast-steel   fittings?     For   low-pressure   cast-iron 
fittings?     In  extra  heavy  cast-iron  fittings? 

7.  How  is  a  lap- weld  made  in  steel  or  iron  pipe?     A  butt-weld? 

8.  For  what  purpose  in  power  plant  steam-piping  may  riveted  pipe  be  used? 

9.  What  per  cent,  of  the  plate-strength  is  secured  with  a  lap- weld?     With  a  butt- 
weld? 

10.  What  is  the  ultimate  strength  of  a  butt-weld  in  a  steel  pipe?     In  a  wrought-iron 
pipe? 

11.  What  is  the  ultimate  strength  of  a  lap-weld  in  a  steel  pipe?     In  a  wrought-iron 


pipe? 

12.  What  is  a  companion-flange? 

13.  How  is  a  companion-flange  shrunk  on  a  pipe-end?     How  welded  on?     How  is  the 
pipe-end  finished  when  the  flange  is  threaded  on?     What  kind  of  a  fit  does  the  flange 
make  with  a  flared  or  lapped  pipe-end? 

14.  What  are  the  purposes  of  pipe-bends?     What  is  the  minimum  advisable  radius  for 
a  pipe-bend?     The  minimum  advisable  tangent-length  for  a  9-in.  pipe  bend  with  shrunk 
flanges? 


384  STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 

15.  Which  is  a  tangent-length  in  a  pipe-bend? 

16.  Enumerate  the  principal  methods  of  distributing  the  steam-output  of  a  set  of 
boilers. 

17.  What  advantage  is  secured  with  duplicate  main  headers? 

18.  What  are  the   main  features  of  the  unit-group  system   of  steam  distribution? 
Why  are  receiver-separators  particularly  necessary  in  the  branch  pipes  to  engines  where 
this  system  is  used? 

19.  What  is  the  commonly-assumed  rate  of  steam-flow  for  live-steam  piping?     For 
exhaust-steam  piping? 

20.  What   is   the   commonly-assumed   range   of  pressure-drop  for  live-steam  piping? 
For  exhaust-steam  piping? 

21.  Describe    a    slip    expansion-joint.     A    swivel    expansion-joint.     A    corrugated 
expansio  n-j  oint. 

22.  How  may  transmission  of  pipe-vibration  be  prevented?     Describe  a  method  of 
support  and  of  suspension  to  localize  pipe-vibration. 

23.  Enumerate  the  common  methods  of  staying  and  supporting  steam-piping.     Enum- 
erate the  special  adaptations  of  each. 

24.  What  is  the  average  percentage  of  heat-saving  effected  with  pipe  coverings? 

25.  Explain  the  operation  of  the  Holly  steam-loop. 

PROBLEMS   ON   DIVISION   11 

1.  The  required  maximum  steam-output  of  a  boiler  is  30,000  Ib.  per  hr.  at  150  Ib 
pressure,  gage.     The  total  length  of  pipe  in  the  lead  to  the  main  header  being  40  ft., 
what  should  be  the  pipe-size? 

2.  Assuming     a     uniform     velocity     of     flow     in    the    main  and  branches,  what 
should  be  the  size  of  a  main  to  supply  four   branches  of  sizes  2.5-in.,  4-in.,  5-in.,  and 
7-in.,  respectively? 

3.  A  6-in.  run  of  steam-pipe  contains  two  globe-valves  and  one  standard  90-deg. 
elbow.     What  length  of  6-in.  pipe  would  offer  equivalent  resistance  to  the  steam-current? 

4.  What   minimum  length  of  pipe  is  permissible  in  making  an  expansion   U-bend  to 
be  used  in  an  8-in.  steam-line,  150  ft.  long,  carrying  steam  at  135  Ib.  pressure  per  sq.  in., 
gage?     The  temperature  of  the  piping,  when  erected,  is  assumed  to  be  60  deg.  fahr. 

5.  What  will  be  the  quantity  of  condensation  in  30  ft.  of  bare  10-in.  steam-pipe  in  an 
atmospheric  temperature  of  90-deg.  fahr.,  if  the  steam-pressure  is  125  Ib.  per  sq.  in., 
gage?     The  external  surface  area  of  10-in.  pipe  is  2.816  sq.  ft.  per  ft.  of  length. 


DIVISION  12 


LIVE-STEAM     AND     EXHAUST-STEAM    SEPARATORS 


455.  A  Live-Steam  Separator  (Fig.  369)  is  a  device  for  re- 
moving entrained  water  from  the  steam  which  is  conveyed, 
through  pipe-lines,  from  boilers  to  various  steam-consuming 
apparatus,  as  reciprocating  engines  and  turbines. 


O 


Pipeline 

from  Boiler-,  .  \ 

Exhaust-; 

/J     .       H-''        Head--' 


To  Feed-Wafer 
Heater., 


Live-Steam 
Separator^ 


::^v:-:viv;^«v^^-;^^^v^^:-,4.i^ 

Fia.  369. — Live-Steam   And   Exhaust-Steam   Separators   Installed   In   Engine   Piping. 

NOTE. — ORDINARILY,  THE  STEAM  ISSUING  FROM  A  BOILER  WHICH  Is 
UNPROVIDED  WITH  SUPERHEATING  SURFACE  MAY  CONTAIN  FROM  0.3 
PER  CENT.  To  5  PER  CENT.  OF  MOISTURE. — If  the  steam  space  of  the 
boiler  is  unduly  restricted,  as  where  an  excessively  large  number  of  tubes 
are  used  in  a  return-tubular  boiler,  the  percentage  entrainment  may 
exceed  greatly  the  maximum  figure  noted  above.  Similarly,  if  a  properly- 
•25  385 


386  STEAM  POWER  PLANT  AUXILIARIES         [Div.  11 

proportioned  boiler  is  forced  much  beyond  its  rated  capacity,  the  entrain- 
ment  may  become  dangerously  excessive. 

NOTE. — MOISTURE  MAY  BE  CARRIED  FROM  A  BOILER  EITHER  As 
FINELY  DIVIDED  SPRAY  OR  As  CONCENTRATED  BULKS  OF  WATER.  It 
may  also  be  due,  wholly  or  in  part,  to  condensation  in  the  pipe-line. 
The  quantity  so  produced  will  depend  largely  upon  the  length  of  the 
pipe  and  the  effectiveness  of  the  covering.  Water  resulting  from  con- 
densation may  accumulate  in  pockets  in  the  piping,  whence  it  may  be 
picked  up  in  bulk  by  the  onrushing  current  of  steam .  Similarly  quantities 
of  water  in  bulk,  or  slugs  of  water,  may  be  projected  from  the  boiler  by 
the  violent  priming  that  may  result  from  a  suddenly  applied  overload, 
or  from  carrying  the  water  too  high  in  the  boiler. 

456.  The  Purposes  Of  Live-Steam  Separation  are:  (1)  To 
conserve  the  energy  of  the  steam.     (2)   To  prevent  wrecking  of 
engines  by  slugs  of  water  which  might  be  present  with  the  steam- 
supply.     (3)    To  prevent  impairment  of  engine-lubrication  by 
wet  steam.     (4)  To  protect  the  valves,  pistons  and  cylinders  of 
reciprocating-engines,  and  the  blades  and  buckets  of  turbine, 
from  the  erosive  action  of  wet  steam. 

NOTE. — MOISTURE  DIMINISHES  THE  NET  THERMAL  VALUE  OF  THE 
STEAM  WHICH  Is  DELIVERED  To  AN  ENGINE,  and,  therefore,  the  thermal 
efficiency  of  the  engine.  It  does  this  by  adding  to  the  initial  condensation 
in  the  cylinder  and  by  absorbing  whatever  superheat  may  be  available 
from  expansion.  A  discussion  of  this  subject  is  contained  in  the  Author's 
PRACTICAL  HEAT. 

NOTE. — ADMISSION  OF  AN  OTHERWISE  TRIFLING  BULK  OF  WATER 
To  AN  ENGINE  CYLINDER  Is  EXTREMELY  DANGEROUS  if  the  engine  is 
running  at  high  speed.  This  is  due  both  to  the  very  restricted  clearance 
spaces  which  considerations  of  economy  demand  for  high-speed  recipro- 
cating engines  and  to  the  fact  that  water  is  practically  incompressible. 

NOTE. — STEAM  TURBINES  MAY  BE  SERIOUSLY  DAMAGED  BY  SLUGS 
OF  WATER  ENTERING  WITH  THE  STEAM.  The  blades  and  buckets  of 
turbines  are  liable  to  be  stripped  by  smaller  masses  of  water  than  such 
as  might  be  required  to  wreck  the  cylinders  of  reciprocating  engines. 

NOTE. — THOROUGH  LUBRICATION  OF  AN  ENGINE-CYLINDER  Is  PRAC- 
TICALLY IMPOSSIBLE  WHEN  EXCESSIVELY  WET  STEAM  Is  USED.  The 
water  will  gather  on  the  rubbing  surfaces  and  thus  exclude  the  oil. 
Otherwise  it  will  precipitate  the  oil  and  flush  it  out  before  it  can  reach 
the  rubbing  surfaces. 

457.  The  Economy  Of  Live -Steam  Separation  is,  aside  from 
the    considerations   previously   noted    (Sec.   456),   mainly   a 
question  of  fuel  saving  which  results  from  delivering  dry  steam 
to  the  prime  mover.     The  loss  from  initial  condensation,  due 


SEC.  458] 


STEAM  SEPARATORS 


387 


to  the  effect  of  wet  steam  in  the  engine  cylinders,  may  be 
regarded  as  approximately  1  per  cent,  for  each  1  per  cent,  of 
moisture  in  the  steam  (Direct  Separator  Company,  STEAM  AND 
OIL  SEPARATORS).  It  may  also  be  assumed  for  turbines  that 
for  each  1  per  cent,  of  moisture  in  the  steam  supplied  there  is 
an  increase  of  about  2  per  cent,  in  the  water  rate  (Harrison 
Safety  Boiler  Works,  SEPARATORS). 

NOTE. — THE  Loss  OF  EFFICIENCY  DUE  To  WET  STEAM  IN  TURBINE 
OPERATION  may  be  ascribed  to  the  extra  friction  which  the  moisture 
creates  within  the  turbine.  The  added  friction  apparently  necessitates 
supplying  an  extra  pound  of  steam  for  each  pound  of  moisture  in  order 
to  maintain  a  proper  velocity  of  flow. 

EXAMPLE. — Assuming  that  10  tons  of  coal,  at  3  dollars  per  ton,  are 
consumed  per  day  in  firing  a  power  plant,  the  saving  which  might  be 
effected  by  a  2  per  cent,  reduction  in  the  moisture  content  of  the  steam 
delivered  to  the  engine  would  annually  amount  to  10  X  3  X  365'  X 
0.02  =  $219. 

458.  The   Principal  Operative  And   Structural  Requisites 
Of  A  Live -Steam  Separator  in  the  supply-line  to  an  engine 
are:  (1)  It  should  afford  the  max- 
imum   attainable  effectiveness  of 

separation.  The  separation 
should  be  (Table  474),  prac- 
tically, 100  per  cent,  effective 
when  the  moisture  entrained 
with  the  steam  is  less  than  5 
per  cent.  It  should  be  at  least 
98  per  cent,  effective  when  the 
entrainment  amounts  to  about 
20  per  cent.  (2)  Its  tendency 
to  reduce  the  pressure  of  the  steam 
should  be  practically  inappre- 
ciable. (3)  It  should  have  storage 
capacity  equal  to  about  four 
times  the  volume  of  the  engine  FIG.  syo.-Verticai  Sections  Through 

/\T  f     Cochrane      Horizontal      Receiver-Sep- 

cyhnder.     (4)    It    should    be    of   arator. 
simple  and  durable  construction. 

459.  A  Live  Steam-Separator  Is  Called  A  Receiver-Separ- 
ator When  It  Is  Provided  With  A  Relatively-Large  Well 

(Fig.  370).     The  well  serves,  both  as  a  receptacle  for  the  water 


< -Dram -j_ 

I-Hcilf  End  Section  I-Siote  Section 


388  STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 

which  is  extracted  from  the  steam,  and  as  a  reservoir  wherein 
an  ample  volume  of  steam  (Sec.  458)  may  be  continuously 
maintained  while  the  engine  is  running. 

460.  The  Steam-Storage  Capacity  Afforded  By  A  Receiver- 
Separator    is    of    three-fold    importance:  (1)  It    operates    to 
prevent  the  vibration  to  which  a  long,  sinuous,  high-pressure 
steam-line  might,  otherwise,  be  liable.     A  prevalent  cause  of 
vibration  of  steam-supply  lines  to  engines  is  the  reaction  which 
results  from  the  sudden  arrest,  at  cut-off,  of  the  steam-current, 
and  the  consequent  impact  of  the  steam  with  the  back  of  the 
valve.     The   constant  volume   of  steam,   which  a  receiver- 
separator  may  maintain  in  close  proximity  to  an  engine  cy- 
linder, acts  as  a  buffer  to  absorb  the  shock  of  such  reaction. 
(2)  It  tends  to  prevent  a  drop  of  pressure  between  the  boiler  and 
the  engine.     The  pressure-drop  in  a  steam-supply  line  may,  in 
the  absence  of  storage  space  close  to  the  engine  cylinder, 
amount  to  10  per  cent,  of  the  boiler  pressure.     (3)  It  acts  to 
prevent  the  excessive  priming  which  might,  otherwise,  attend  a 
suddenly  applied  overload. 

461.  An  Exhaust-Steam  Separator  (Fig.  369)  is  a  device 
for  removing  oil  from  the  steam  which  has  been  used  in  engine- 
cylinders  and  expelled  therefrom. 

NOTE. — Exhaust-steam  separators  are  commonly  called  oil-separators 
and  oil-eliminators. 

462.  The  Main  Purposes  Of  Exhaust-Steam  Separation 

are:  (1)  To  render  the  steam  suitable  for  use  in  open  feed-water 
heaters  and  thereby  conserve  the  heat  therein.  Oil  in  the 
feed-water  is  dangerous  to  the  integrity  of  steam-boilers. 
(2)  To  preserve  the  radiating-effectiveness  of  exhaust-steam 
heating  systems.  (3)  To  preserve  the  condensing-effectiveness 
of  surface  condensers.  A  film  of  oil  in  the  radiators  of  a  heating 
system,  or  in  the  tubes  of  a  condenser,  greatly  retards  the 
transmission  of  heat  from  the  steam  to  the  external  air  in  the 
one  case,  or  to  the  cooling  water  in  the  other.  (4)  To  render 
available,  for  boiler  feed^water,  the  discharge  from  surface 
condensers. 

463.  The  Economy  Of  Exhaust-Steam  Separation  is,  aside 
from    the    principal    considerations    previously    enumerated 


SEC.  464]  STEAM  SEPARATORS  389 

(Sec.  462),  largely  a  question  of  the  saving  which  purification 
of  the  exhaust-steam  effects  in  the  cost  of  water  for  operating 
the  plant.  Where  the  boiler  feed-water  is  taken,  without 
cost,  from  streams  or  other  nearby  sources,  conservation  of  the 
water  supply  is  of  little  moment.  But  where  the  boiler-water 
is  taken  from  city  mains,  the  expense  of  wasting  the  exhaust- 
water  may  assume  serious  proportions. 

EXAMPLE. — Allowing  14  pounds  of  feed  water  per  hour  per  horsepower 
developed  by  a  set  of  condensing  engines  the  annual  water-consumption 
for  this  purpose  would  be  about  14  Ib.  X  24  hr.  X  365  days  -^  62.5  Ib. 
per  cu.  ft.  =  1,962  cu.  ft.  per  h.p.  If  the  water  costs  $0.50  per  1,000 
cu.  ft.,  and  the  plant  develops  a  daily  average  of  10,000  h.p.,  the  annual 
expense  for  boiler-feed,  if  the  discharge  from  the  condensers  were  wasted, 
would,  therefore,  be  (1,962  X  10,000  -r-  1,000)  X  0.50  =  $9810.00.  As- 
suming, in  this  case,  that  80  per  cent,  of  the  condensed  exhaust  steam 
were  returned  to  the  boilers  as  clean  feed-water,  the  annual  saving  would 
be  9,810  X  0.8  =  $7848.00. 

464.  The  Physical  Phenomena  Involved  In  The  Operation 
Of  Steam-Separators  are:  (1)   Expansion.     (2)   Momentum. 
(3)  Elasticity.     (4)    Capillary  entrainment.    '(5)    Absorption. 
These    principles,    as    explained    hereinafter,    are    variously 
applied.     The  first  four  are  observable  in  the  operation  of  all 
separators. 

NOTE. — EXPANSION,  As  A  PRINCIPLE  OF  SEPARATION,  is  prominent  in 
the  workings  of  all  receiver-separators.  The  current  of  steam  expands 
somewhat  after  issuing  from  the  contracted  pipe  passage  (P,  Fig.  370) 
into  the  relatively-ample  volume  of  the  receiver,  R.  Its  density  thus 
momentarily  diminishes.  Hence,  it  becomes  less  effective  for  supporting 
the  suspended  moisture.  The  tendency  of  the  water  particles  to  drop 
out  of  the  steam  by  their  own  weight  is,  therefore,  increased. 

465.  Steam  Separators  May  Be  Classified  According   To 
Their  Principal  Modes  Of  Operation  as  follows:  (1)  Reverse- 
current  separators.     (2)  Centrifugal  separators.     (3)  Impact  or 
Baffle-plate   separators.     (4)    Mesh   separators.     (5)    Gridiron 
separators.     (6)  Absorption  separators. 

466.  The   Main  Operating  Principle   Of  Re  verse -Current 
Separators  (Figs.  371,  372  and  373)  is  the  momentum  which  a 
body  acquires  through  propulsion  by  a  force  acting  along  an 
approximately   straight  line.     After  entering  the  separator, 


390 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 


the  moisture-laden  current  of  steam  traverses  a  short  distance 
(Fig.  371)  in  a  direct  line.  Its  course  is  then  reversed  abruptly. 
The  steam  readily  adjusts  itself  to  the  altered  direction  of 
flow.  But  the  water  particles  being  of  much  greater  specific 
gravity  than  the  steam,  are  propelled  by  their  own  momentum 
to  the  bottom  of  the  separating  chamber. 

NOTE. — With  the  separation  shown  in  Fig.  371,  removal  of  the  mois- 
ture depends  solely  upon  the  whip-snap  action  which  accompanies  the 
current-reversal.  With  the  apparatus  shown  in  Fig.  372,  two  horizontal 
baffle-plates  or  wings,  one  projecting  laterally  from  each  side  of  the 


Water-Gage- 


Drainage  Ducts  Leading 
From  Troughs  To  Bottom 
Of  Separating  Chamber 

Fio.  371.— Happes  Re-  Fio.  372.— Welderon  Re- 
verse-Current Horizontal  verse-Current  Horizontal 
Exhaust-Steam  Separator.  Receiver-Separator. 


Drain-- 

Outlet--* 'Cutlet  Tube 
Hooded  Guard  to  Prevent-- 
Upward Creepage  of  Water 


Fio.  373.— Austin  Re- 
verse-Current Vertical 
Live-Steam  Separator. 


diaphragmed  steam-duct,  aid  in  the  separation.  With  the  apparatus 
shown  in  Fig.  373,  the  separation  is  partially  effected  by  impact  of  the 
current  with  the  hoods. 

467.  The  Main  Operating  Principle  Of  Centrifugal  Sepa- 
rators (Figs.  374,  375,  376)  is  the  tangential  momentum  which 
a  body  acquires  through  the  action  of  centrifugal  force.  The 
steam-current  assumes  a  spiral  or  twisting  motion  at  the 
instant  of  its  entrance  to  the  separator.  The  centrifugal 
force  thereby  developed  in  the  particles  of  oil  or  water  impels 
them  to  fly  tangentially  from  the  steam-current.  Thus,  the 
oil  or  water  is  flung  against  the  inner  surface  of  the  external 
shell,  down  which  it  trickles  to  the  drainage  outlet. 


SEC.  467] 


STEAM  SEPARATORS 


391 


NOTE. — The  device  for  imparting  a  twisting  motion  to  the  steam  in  a 
centrifugal  separator  may  be  a  helix  in  the  throat  of  the  inlet  orifice 


Steam  In/e 


.-Helix  for  Throwing*    Slot  with  Over- 
'       OilorWb/fer        '-. stopping  Edges 


••••Drain i 

I -End  Section          n-Side  Section 

FIG.  374. — Swartwout  Centrifugal        FIG.  37.5. — Masher  Centrifugal  Horizontal  Steam 
Steam  Separator.  Separator. 


Inlet- 


Fia.  376. — Stratton  Centrifugal  Hori- 
zontal Live-Steam  Separator. 


(Primary  Separation  by  !m- 

'. pact  or  Current  with  Baffle 

Irilet-.^ 


'Secondary  Separation  by 
Reversal  of  Current 

FIG.  377.— Austin  Baffle-Plate  Angle 
Live-Steam  Separator. 


(Fig.  374),  a  helix  which  traverses  the  interior  of  the  separator  from  the 
inlet  to  the  outlet  (H,  Fig.  375),  or  a  spiral  web  (W,  Fig.  376)  which 
winds  about  a  central  outlet  tube. 


392 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 


468.  The  Main  Operating  Principle  Of  Impact  Or  Baffle- 
Plate  Separators  (Fig.  370,  377,  378,  379,  380  and  381)  is  the 


IWatir 


'Sectncfotry  Separa 
t Jon  by  Reversal  of  ";' 
Steam-Current---'' 


•Dram 


FIG.  378.—"  Austin  "  Baffle-Plate  Under- 
slot  Horizontal  Live-Steam  Separator. 


.  •  •  -Ribbea/  Baffles-  - 
:Stzam  Vut/et  Port 


.-Horizontal  Section 

Through  Baf f \&5  omci  P?rrj 

-Glass  Water-Gage 
••-Water  Chamber 

.  — Drain 

I-Pcrspectiv<TSriowincf 
Baffles 

FIG.   379.— "Baum"  Baffle-Plate  Hori- 
zontal-Steam Separator. 


ttid-Water  Sprays 

*»  £\  _  ,  -- 


.--Cold-Water  Spray 
Pipe. 

..-Inlet 


.-Flange-Gutter 

'for  Catching  Oil 

which  Gathers  in 

Exhaust  Pipe 

.-Drainage 
•      Hole 


Connection  fo 
~- Auxiliary  Vacuum 


Pump 


FIG.  380. — Austin  Baffle-Plate  Horizon- 
tal Exhaust-Steam  Separator  For  Vacuum 
Service. 


Ribbed  Baffle  Formed  by 
Wall  of  Cold  Water  Circulating 


Equalizing-' 
Pipe  Connection 


FIG.  381. — Baum  Baffle-Plate 
Horizontal  Exhaust-Steam  Sep- 
arator For  Vacuum  Service. 


elasticity  of  steam.  The  entering  steam-current  (Fig.  377) 
impinges  upon  the  upper  baffle,  B.  Due  to  its  great  elasticity, 
the  steam  rebounds  therefrom.  But  the  quite  inelastic  water 


SEC.  469] 


STEAM  SEPARATORS 


393 


•Inlet  .-.-Sieve 

/.•Centering.  W/ngs 


adheres  to  the  plate  and  trickles  by  capillary  entrainment  into 
the  trough  at  its  lower  edge.  Thence  it  flows  to  the  drainage 
outlet,  O.  The  separation  thus  far  is,  however,  only  partial. 
When  the  steam  rebounds  the  upper  baffle,  it  strikes  the  outer 
shell,  S.  It  then  rebounds  downward,  toward  the  opening  to 
the  lower  baffle,  and  reverses  its  direction  of  flow.  Additional 
moisture  is  thus  whipped  out  by  its  own  momentum. 

469.  Corrugated  And  Fluted  or  Ribbed  Surfaces  In  Steam 
Separators  (Figs.  370,  378,  379,  380  and  381)  perform  a  two- 
fold function:  (1)  They  prevent  the  sweep  of  the  steam-current 
from  scouring  the  adhering  particles  of  oil  or  moisture  from  the 
surfaces.     (2)  They  facilitate  the  tendency  of  the  separated  oil 
or  water  to  trickle  downward  in  a  multitude  of  small  individual 
streams. 

470.  The  Main  Operating  Principle  Of  Mesh  Separators 
(Fig.  382)  is  the  tendency  of  fluid  particles  to  entrain  and 
form  into  minute  rivulets 

.  .  „  Primary  Separation         „  -  -Conical  Top  of  Hmt 

by  capillary  attraction.  ' 
The  entering  steam-cur- 
rent, E,  impinges  directly 
upon  the  sieve,  S,  which 
covers  the  conical  top  of 
the  hood,  H,  surrounding 
the  upper  orifice  of  the 
outlet  tube,  0.  A  portion 
of  the  water  or  oil  will  ad- 
here to  the  sieve,  and,  by 
capillary  entrainment,  will 
pass  through  its  meshes 
to  the  top  surface  of  the 
hood.  The  water  or  oil  thus 
deposited  flows  through  the 
drainage  tubes  D,  to  the 
collecting-chamber,  C. 

Impact  with  the  conical  surface  changes  the  form  of  the 
steam-current  to  that  of  an  annular  sheet  which  sweeps  down- 
ward in  the  space  between  the  cylindrical  wall  of  the  hood,  H, 
and  the  cylindrical  sieve  or  trapping-sheet,  T.  Nearly  all  of 
the  remaining  moisture,  or  oil,  is  caught  in  the  meshes  (Fig. 


Secondary  Sep- 
aration by  Rever- 
sal of  Current  ' 


Drain- 


Outlet- 


FIG.  382. — Sweet   Mesh   Vertical  Steam  Sep- 
arator. 


394 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 


383)  of  the  trapping-sheet,  T.  It  is  thereby  entrained  in  tiny 
streams  which  flow  to  the  annular  space,  A,  between  the  trap- 
ping sheet  and  the  shell.  Thence  it  trickles  downward  to  the 


;Sheet-Stee! 


Side  Views  of  Hol- 
low Grid-Columns- - 


Outlet-.. 


Vertical  Sections 
of  Hollow  Grief  - 
Columns-  ~. 


External-1 
Shell 


/ar  Space 
',- between  Trap - 
\pinpj  Sheet  and 

'"A  '  5he' 


FIG.  383. — Sectional  Detail  Of 
Sweet  Steam  Separator. 


Oil  or  Water  Dropping 
from  Channels  in 
Snot -Columns 


••Drain 


FIG.    384. — Bundy    Gridiron    Horizontal    Steam 
Separator. 


to  Hollow  Interiors 
of Columns-^ 


collecting  chamber,  C.  Practically  all  of  the  moisture,  or  oil, 
which  still  remains  in  the  steam-current  will  be  whipped  out 
as  the  current  reverses  its  direction  of  flow  in  passing  upward 
to  the  outlet-tube  orifice,  0.  The  per- 
forated diaphragm,  P,  prevents  the  steam- 
current  from  picking  the  water,  or  oil  and 
water,  out  of  the  chamber  beneath. 

471.  The  Main  Operating  Principle  Of 
Gridiron  Separators  (Fig.  384)  is  capillary 
attraction.  A  series  of  gridiron  separat- 
ing-plates  (Fig.  385)  is  arranged  in  stag- 
gered formation  (Fig.  386)  in  the  path  of 
the  steam-current.  The  columns  of  these 
plates  are  hollow.  Vertical  series  of 
small  cups,  or  recesses,  are  cast  in  the 
faces  of  the  columns  against  which  the 
entering  steam  impinges.  A  small  hole  is  drilled  from  each 
cup  to  the  hollow  interior  of  the  column.  The  particles  of 


fi 
M 


Recesses  In  : 


FIG.  385.  —  Gridiron 
Separating-P  late  Of 
Bundy  Steam  Separator. 


SEC.  472] 


STEAM  SEPARATORS 


395 


water,  or  oil,  are  projected  against  the  grids  and  cling  thereto. 
The  capillary  action  which  then  ensues  causes  them  to  gather 
in  the  cups.  Thence  they  trickle  through  the  small  ports 
which  lead  to  the  channels  inside  the  columns  From  these 
they  fall  into  the  collecting-chamber,  C,  beneath. 

472.  The  Operating  Principle  Of  Absorption  Separators 
(Fig.  387)  depends  upon  the  absorbent  properties  of  certain 
porous  or  fibrous  materials. 


s-rDrilleol  Ports,  Leading 
:  from  Recesses  to'  Hoi  low 
[  Interiors  of  Grio/-Co/umn 


Ano/te-Plate-. 
' 


Outlet-. 


FIG.  386.  —  Staggered  Formation  Of 
Gridiron,  Separating  Plates  In  Bundy 
Steam  Separator. 


Ducts 


FIG.  387. — Loew  Absorption  Exhaust 
Steam  Separator. 


NOTE. — Absorption  separators  are  designed  only  for  exhaust-steam 
separation. 

473.  The  Maximum  Efficiency  Of  Separation  Attainable 

(Table  474)  with  any  given  type  of  live-steam  separator 
varies  according  to  the  quality  of  the  steam  as  it  enters  the 
separator.  (See  Sec.  476  for  meaning  of  efficiency.) 

NOTE. — The  efficiency  of  a  live-steam  separator,  and,  therefore,  the 
ultimate  effectiveness  of  separation  is  benefited  by  providing  an  ade- 
quate covering  of  insulating  material. 


396 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 


474.  Table   Showing   Efficiencies   Obtained   In  Tests    Of 
Live-Steam  Separators  Of  Six  Different  Makes. 

(From  Power,  May  11,  1909) 


Steam  with  less 

Steam  with  about 

Steam  with  about 

than  5  per  cent,  of 

10  per  cent,  of 

20  per  cent,  of 

moisture 

moisture 

moisture 

Make 

of 
sep- 

Quality 
of  steam 

Qualityi 
of  steam 

Quality 
of  steam 

Qualityi 
of  steam 

Quality 
of  steam 

Qualityi 
of  steam 

Efficiency, 
per  cent. 

arator 

before 

after 

before 

after 

before 

after 

separa- 

separa- 

separa- 

separa- 

separa- 

separa- 

tion 

tion 

tion 

tion 

tion 

tion 

A 

97.5 

99.0 

60.0 

87.0 

98.8 

.... 

90.8 

78.1 

98.8 

94.5 

B 

96.1 

97.4 

33.3 

- 

90.1 

98.0 

80.0 

79.5 

98.2 

91.2 

C 

98.1 

98.5 

21.1 

89.6 

95.8 

59.6 

81.7- 

'-    97.9 

83.5 

D 

97.7 

97.9 

8.7 

90.6 

93.7 

33.0 

78.2 

95.6 

79.8 

E 

95.6 

95.8 

4.5 

88.9 

92.1 

28.8 

82.4 

90.4 

45.5 

F 

98.0 

98.0 

0.0 

88.4 

90.2 

15.  5 

79.3 

87.2 

38.1 

1  Denotes  effectiveness  of  separation. 

475.  The  Velocity  Of  The  Steam-Current  In  Transit 
Through  A  Separator  Affects  The  Efficiency  Of  The  Separator. 
The  efficiency  diminishes  as  the  velocity  increases.  If  a  sepa- 
rator is  so  designed  as  to  permit  an  excessive  velocity  of 
steam-flow  through  it,  its  efficiency  (Fig.  388)  may  be  practi- 
cally zero. 


NOTE. — Expansion  of  the  steam  (Sec.  460)  in  transit  through  the  rela- 
tively-large steam  space  of  a  separator  results  in  a  momentary  diminution 


SEC.  476] 


STEAM  SEPARATORS 


397 


of  the  velocity  of  flow.  The  initial  velocity  is,  however,  restored  when 
the  steam  reenters  the  outlet  pipe  if  the  outlet  is  the  same  size  as  the 
inlet  pipe. 


Average  Initial — 
Quctljty  ofSteam=_ 
"'90%'" 


2000  3000  4000          £000  6000  7000  8000 

Velocity    of     Steam-    Feet     Per      Minute 

Fia.  388.— Graph  Showing  Relation  Between  Efficiency  Of  Separation  And  Velocity  O 

Steam  Flow. 

476.  The  Efficiency  Of  A  Live-Steam  Separator  may  be 

computed  by  the  following  formula : 

100  W 
(105)  E  =  — == — w    (per    cent,   efficiency) 

Wherein  E  =  per  cent,  efficiency.     Ww  =  weight  of  separated 


Fia.   389. — Arrangement  Of  Separator  And  Appurtenances  For  Efficiency  Test. 

water,  in  Ib.  W*  =  weight  of  moisture,  in  lb.,  in  a  definite 
weight  of  steam  delivered  to  the  separator,  as  determined 
(Fig.  389)  by  calorimeter  and  steam-flow  tests. 


398  STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 

EXAMPLE.  —  A  steam-flow  meter  at  S  (Fig.  389),  records  a  flow  of 
16,273  Ib.  of  steam  during  a  certain  time-interval.  A  calorimeter  at  C 
shows  the  quality  of  the  steam  to  be  94.5  per  cent.  The  weight  of  the 
separated  water  drawn  during  the  interval  from  the  storage  reservoir,  R, 
is  530  Ib.  What  is  the  efficiency  of  the  separator?  SOLUTION.  —  The 
weight  of  moisture  in  the  steam  =  16,273  X  (1  -  0.945)  =  895  Ib. 
Applying  For.  (105),  E  =  WQWw/Wi  =  100  X  530  -f-  895  =  59.2  per 
cent. 

476A.  The  Efficiency  Of  A  Live  Steam  Separator  May 
Also  Be  Computed  On  The  Basis  Of  The  Quality  Of  The 
Steam  Entering  And  Leaving  The  Separator  by  applying  the 
following  formula: 


(105A)  E=  (per  cent.  efficiency) 

luu  —  x\ 

Wherein:  x\  —  quality  of  the  steam  entering  the  separator, 
in  per  cent.  z2  =  quality  of  the  steam  leaving  the  separator, 
in  per  cent. 

EXAMPLE.  —  In  the  preceeding  example,  what  is  the  efficiency  of  the 
separator  if  a  calorimeter  at  B  (Fig.  389)  shows  the  quality  of  the  steam 
leaving  the  separator  to  be  97.8  per  cent.? 

SOLUTION.—  By  For.  (105A):  E  =  100(z2  -  Zi)/(100  -  zi)  =  100  X 
(97.8  -  94.5)  -T-  (100  -  94.5)  =  59.2  per  cent. 

477.  Exhaust-Steam  Separation  In  A  Partial  Vacuum  May 
Be  Facilitated  By  Wetting  The  Separating  Surface.—  This 
may  be  accomplished  in  either  of  two  ways:  (1)  By  injecting 
(Fig.  380)  a  spray  of  cold  water  against  the  surface.  (2)  By 
circulating  cold  water  within  a  chamber  (Fig.  381)  the  wall  of 
which  forms  the  separating  surface.  Moisture  is  thus  by 
condensation  of  a  portion  of  the  steam,  caused  to  appear  on 
the  separating  surface. 

NOTE.  —  When  an  engine  is  exhausting  into  a  partial  vacuum  the  steam 
will  have  little  tendency  to  condense  or  to  entrain  moisture  during  its 
passage  from  the  engine-cylinder  to  the  condenser.  Hence,  all  of  the 
surfaces  which  the  steam  encounters  will  continue  dry.  The  fine  parti- 
cles of  cylinder  oil  will,  therefore,  due  to  their  very  low  specific  gravity, 
tend  to  rebound  with  the  steam  from  the  separating  surface.  But  if 
the  surface  is  covered  with  a  film  of  moisture,  the  moisture  will  diffuse 
the  oil-particles  over  the  surface  and  thus  cause  them  to  adhere  thereto. 


SEC.  478] 


STEAM  SEPARATORS 


399 


478.  An  Exhaust-Head  (Figs.  390  and  391)  is  an  exhaust- 
steam  separator  especially  designed  for  attachment  to  the 
discharge-end  of  an  engine  exhaust-pipe  which  opens  to  the 
atmosphere. 


Partial  Separation  by  Impact, 
(AcfheiyrKe.anol  Capillary  Entrainrnerrt 

Upper  Baffle- 
Cone 


Cylindrical  Trapping 


.    Partial 
Separation  by 
Current  Reversal 

''-Drain 
Exhaust-Pipe 
Connection-. 


FIG.  390.— "Wright"  Baffle-Plate  Exhaust- 
Head. 


•feme  , 
Trapping 
Sheet- 

Parfia!-'' 

Separation 

byCurrent    ^-Exhaust-Pipe 

•Reversal        Connection 

FIG.     391.— "Sweet"    Mesh    Ex- 
haust-Head. 


479.  The  Purpose  Of  An  Exhaust-Head  Is  twofold:  (1)  To 
prevent  pollution  of  the  atmosphere  and  befoulment  of  the  roofs 
and  walls  of  buildings  by  the  oil-and-water  in  the  exhaust-steam. 
(2)  To  muffle  the  sound  of  the  exhaust. 

480*  The  Proper  Location  For  A  Live -Steam  Separator  is 
as  close  to  the  apparatus  which  it  is  designed  to  serve  as  the 
piping  arrangement  will  permit.  Where  the  separator  is 
used  (Fig.  369)  in  connection  with  an  engine,  it  should  be 
connected  directly  to  the  throttle  valve. 

481.  The  Proper  Location  For  An  Exhaust-Steam  Separator 
depends  upon  the  ultimate  disposition  of  the  exhaust.  In  a 
non-condensing  plant,  the  separator  may  be  installed  (Fig.  369) 
in  the  main  exhaust  pipe  close  to  the  point  where  it  branches 
to  the  feed-water  heater  and  the  radiator  heating  system.  In 
a  surface-condensing  plant,  the  separator  may  be  installed 
at  any  point  between  the  engine  and  condenser.  If  a  vacuum 
feed-water  heater  (Sec.  249)  is  included  in  the  installation, 
and  the  separator  is  unprovided  with  a  device  for  wetting 
the  separating  surfaces,  it  may  be  preferable  to  place  the 


400 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 


separator  between  the  heater  and  condenser.  The  moisture 
which  the  steam  entrains  in  the  heater  will  thus  become 
available  for  wetting  the  surfaces.  A  disadvantage  of  this 
arrangement  is  that  the  heater-tubes  will  be  exposed  to 
befoulment  by  the  oil. 

NOTE. — Exhaust-steam  separators  are  not  commonly  used  in  connec- 
tion with  condensers  in  which  the  steam  mingles  directly  with  the  con- 
densing water. 

482.  The  Selection  Of  A  Suitable  Live-Steam  Separator 

is  mainly  a  question  of  adapting  its  shape  to  structural  limi- 
tations. The  vertical,  horizontal,  and  angle  forms  provide 
flexibility  of  choice  in  this  regard.  Otherwise,  it  is  usually 
only  necessary,  when  ordering  a  separator,  to  specify  the  size 
of  the  steam-pipe,  the  type  of  engine  and  the  steam-pressure. 
The  proportions  adopted  by  the  different  manufacturers  are 
made  conformable  to  these  data. 

NOTE. — The  size  of  a  steam-separator  refers  to  the  size  of  the  pipe-line 
in  which  the  separator  is  installed. 

483.  The  Selection  Of  A  Suitable  Exhaust-Steam  Separator 
is  mainly   contingent   upon   the   following   information:  (1) 

The  number  and  sizes,  of  the 
engines,  including  steam-pumps, 
which  are  to  exhaust  through 
the  separator.  (2)  The  required 
location  of  the  separator  (Sec. 
481).  (3)  Whether  the  plant  is 
operated  condensing  or  non- 
condensing.  (4)  The  pressure  of 
the  exhaust.  (5)  The  quality  and 
quantity  of  the  cylinder  oil  used. 


FIG.  392. — Eclipse  Exhaust  Steam 
Separator  Arranged  To  Reduce  Veloc- 
ity Of  Steam  Flow. 


NOTE. — The  first  and  fourth  items  enumerated  above  mainly  deter- 
mine the  velocity  of  flow  through  the  main  exhaust-pipe.  The  slower 
the  steam-flow,  the  more  effective  the  separation.  Adequate  separation 
may,  therefore,  be  generally  insured  by  selecting  a  separator  (Fig.  392) 
two  or  three  sizes  larger  than  the  exhaust  pipe  size. 

484.  A  Live-Steam  Separator  Should  Be  Drained  Auto- 
matically (Fig.  369)  by  a  reliable  steam  trap.  (See  Div.  13.) 


SEC.  485] 


STEAM  SEPARATORS 


401 


Steam-Sefxxrcttor- 


Draln  Connection-' 


FIG.  393.  —  Device  For  Shielding 
Glass  Gage  From  Fluctuations  Of 
Steam-Temperature. 


485.  Steam-Separators  Should  Be  Equipped  With  Glass 
Water-Gages    (Fig.  369).     The   glass-gage,  G,  may  be   con- 
nected in  parallel  with  a  by-pass  pipe  (P,  Fig.  393).     The  pur- 
pose of  this  arrangement  is  to  minimize  glass  breakage;  See 
Power  1910. 

NOTE. — The  breakage  to  which  glass  gages  are  peculiarly  susceptible 
when  attached  to  steam  separators  may  be  due  to  the  frequent  and  rapid 
changes  of  temperature  to  which  the  glass  is  subjected.  The  pressure 
within  a  separator  in  the  supply  pipe  of 
an  engine  may  fluctuate  through  a 
range  of  perhaps  10  pounds.  This  will 
be  accompanied  by  a  fluctuation  in 
temperature  which  may  affect  the 
molecular  structure  of  the  glass.  The 
glass  will  crystallize  quickly  and  will 
eventually  shatter  into  fragments. 
By  locating  the  gage  at  a  considerable 
distance  from  the  separator  and  in- 
troducing an  intermediary  passage 
(P,  Fig.  393),  sufficient  condensation 
may  be  thereby  induced  to  cause  a 
thin  film  of  water  to  gather  on  the  interior  of  the  glass.  This  moisture 
will  diminish  -by  evaporation  as  the  pressure  drops  and  will  augment  by 
further  condensation  as  the  pressure  rises.  Thus  it  may  minimize 
temperature  fluctuation  in  the  glass. 

486.  The   Cost  Of  Steam  And  Oil  Separators:  Standard 
horizontal-type  oil  separators,  2  to  8  in.,  range  in  price  $8  to 
$36.     Vertical-receiver-type  oil  separators,  2  to  8  in.,  $13.60 
to  $62.00.     Standard  vertical  steam  separators,  2  to  8  in., 
$18.40    to   $88.00.     Standard    horizontal   steam  separators, 
2  to  8  in.,  $12  to  $52.     Preceding  values  (from  MECHANICAL 
AND  ELECTRICAL  COST  DATA,  Gillette  and  Dana,  McGraw- 
Hill)   are  pre-war  costs.     During  and  immediately  after  the 
great  war  the  prices  were  advanced  from  about  100  per  cent, 
for  the  small  to  25  per  cent,  for  the  large  sizes. 

QUESTIONS  ON  DIVISION  12 

1.  What  is  a  live-steam  separator? 

2.  What  percentage  of  entrained  moisture  does  the  steam  delivered  by  a  boiler,  with- 
out superheating  surface,  ordinarily  contain? 

3.  What  circumstances  of  boiler-design  and  operation  principally  affect  the  degree  of 
moisture-entrainment? 

4.  What  are  slugs  of  water  in  a  steam  pipe?     What  causes  the  entrained  moisture  to 
form  slugs? 

26 


402  STEAM  POWER  PLANT  AUXILIARIES         [Div.  12 

5.  What  contributory  circumstance  usually  determines  the  total  quantity  of  moisture 
in  the  steam  delivered  to  a  separator? 

6.  Enumerate  the  chief  purposes  of  live-steam  separation. 

7.  Through  what  phenomena,  occurring  within  an  engine  cylinder,  is  diminishment 
of  the  engine's  thermal  efficiency  by  wet  steam  mainly  effected? 

8.  Why  are  slugs  of  water  in  the  steam-supply  particularly  dangerous  to  high  speed 
reciprocating  engines? 

9.  In  what  way  may  damage  occur  to  a  turbine  by  small  masses  of  water  in  the  steam- 
supply? 

10.  How  does  wet  steam  affect  the  internal  lubrication  of  an  engine? 

11.  What  approximate  numerical  relation  exists  between  the  percentage  of  moisture 
in  the  steam  delivered  to  an  engine  and  the  resulting  percentage  of  loss  of  economy? 

12.  What  circumstance  apparently  explains  the  loss  of  thermal  efficiency  that  results 
from  supplying  wet  steam  to  a  turbine? 

13.  What  are  the  chief  requisites  of  a  live-steam  separator? 

14.  What  is  a. receiver-separator? 

15.  What  benefits  may  attend  the  use  of  receiver-separators? 

16.  How  does  a  receiver-separator  operate  to  prevent  vibration  of  the  steam-supply 
pipe  of  an  engine? 

17.  What  is  an  exhaust-steam  separator? 

18.  What  are  the  principal  purposes  of  exhaust-steam  separation? 

19.  What  is  the  outstanding  consideration  with  respect  to  the  economy  of   exhaust- 
steam  separation? 

20.  What  are  the  physical  phenomena  which  are  mainly  observable  in  the  operation 
of  steam  separators? 

21.  How  does  expansion  of  the  steam  affect  separation? 

22.  Enumerate  the  general  classes  of  steam-separators. 

23.  What  is  the  main  operating  principle  of  reverse-current  separators?     Of   centri- 
fugal   separators?     Of    baffle-plate    separators?     Of    mesh    separators?     Of     gridiron 
separators?     Of  absorption  separators? 

24.  What  are  the  functions  of  corrugations  and  ribs  on  the  inner  surfaces  of  steam- 
separators? 

25.  What  variable  factor  controls  the  operating  efficiency  of  a  live-steam  separator? 

26.  What  factors  determine  the  operating  efficiency  of  a  separator?     What    factor 
determines  the  effectiveness  of  the  separation  accomplished  by  a  separator? 

27.  What  effect   will  diminished  velocity  have  on  the  efficiency  of  the  separator? 
How  may  a  diminished  velocity  of  flow  through  a  separator  be  obtained? 

28.  Why  may  advantage  result  from  injecting  water  into  the  exhaust-steam  separator 
of  a  condensing  engine? 

29.  What  is  an  exhaust-head? 

30.  What  are  the  functions  of  an  exhaust-head? 

31.  What  circumstances  mainly  govern  the  selection  of  a  proper  point  of  location  for 
a  steam  separator  in  an  exhaust-line? 

32.  What  considerations  are  principally  involved  in  the  selection  of  a  live-steam 
separator?     Of  an  exhaust-steam  separator? 

33.  What  benefit  may  result  from  installing  an  exhaust-steam  separator  of  larger 
size  than  the  exhaust- pipe  size? 

34.  How  should  live-steam  separators  be  drained? 

35.  To  what  inherent  circumstance  of  operation  may  difficulty  of  maintaining  glass 
water-gages  on  separators  be  ascribed? 

PROBLEMS  ON  DIVISION   12 

1.  In  a  certain  locality,  coal  is  available  at  $4.00  per  ton.     If  30  tons  are  normally 
consumed  per  day,  what  will  be  the  saving  per  year  if  the  quality  of  the  steam  delivered 
to  the  reciprocating  engines  is  raised  by  a  separator  from  95  to  98  per  cent.? 

2.  The  steam  passing  to  a  certain  separator  has  a  quality  of  93  per  cent.    If  5,600  Ib. 
pass  per  hour  and  the  separator  collects  285  Ib.  of  water,  what  is  the  efficiency  of  the 
separator? 


DIVISION  13 
STEAM  TRAPS 

487.  Steam  Traps  are  devices  for  entrapping  and  auto- 
matically disposing  of  the  water  that  results:  (1)  From  con- 
densation and  entrainage  in  steam-piping  systems  (Fig.  394,  395 
and  396).  (2)  From  condensation  in  steam-heating  apparatus, 
(3)  From  condensation  in  steam-power  apparatus  (Figs.  397 
and  398). 

time  Eliminator  or  Live- Steam  Separation, 
Steotmline  to  Engine* 


Corrugated  Baft  let 

s  to  Prevent 
Current  from 
up  Separated 

Water 
Vent-Pipe- 


Fia.  394. — Nason  Bucket-Float  Intermittent-Discharge  Medium-Pressure  Steam  Trap 
Installed  For  Draining  A  Live-Steam  Separator. 

NOTE. — STEAM  TRAPS,  IN  GENERAL,  MAY  BE  DIVIDED  INTO  Two 
GROUPS:  (1)  Return  traps  (Fig.  399)  or  those  which  discharge,  against 
boiler-pressure,  directly  into  the  water  spaces  of  steam  boilers.  (2) 
Non-return  traps  (Fig.  400)  or  those  which  discharge  against  normal 
atmospheric  pressure,  or  into  receptacles  under  less  than  boiler  pressure. 

STEAM  TRAPS  MAY  BE  CLASSIFIED  ACCORDING  To  THE  PRINCIPLES 
Or  OPERATION  CHIEFLY  EMPLOYED  as:  (1)  Buoyancy  traps,  which  com- 
prise ball-float  traps  (Fig.  396,  400  and  401 )  and  bucket-float  traps  (Figs. 
394  and  398).  (2)  Counierweighied  tilting  or  dumping  traps  (Fig.  399). 
(3)  Expansion  traps  (Figs.  395  and  397). 

403 


404 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  13 


-Riser 


Water 
Column — ' 


5 '/ rc/mer-- ';'B 


•Mloy  Expansion  Tubt 


FIG.  395. — Kieley  Expansion  Intermittent-Discharge  Steam  Trap  Draining  Radia- 
tion. When  T  Fills  With  Water  And  Cools,  It  Contracts  And  Draws  In  H  And  P. 
V  Is  Then  Opened  By  Upward  Thrust  Of  S  Against  L.  When  Steam  Enters,  T  Ex- 
pands And  Pushes  Out  77  And  P.  V  Is  Then  Closed  By  Downward  Thrust  Of  P 
Against  L. 


Dry-Vacuum  P/pe---~^  . 


Steam  Supply  for 
Discharge  Against 
Hydrostatic  Head--, 


Ball-Valve  for  Controlling 
Admission  ofAiror-Sfearr. 
Pressure  for  Discharge- 


Ball-Valve  for  Controlling 
Connection  between  Trap 
and  Vacuum  System-^ 


Glass 
Water-Gauge--'' 


Lifting  Discharge 
Pipe- >- 


Gravity-Discharge  Pipe- ,  ^ 


Fio.  396. — Strong  Vacuum  Trap  Installed  For  Draining  Separator  In  Condensing- 
Engine  Exhaust-Line.  When  F  Rises,  V  Closes  And  P  Opens,  Permitting  Live  Steam 
Or  Atmospheric  Air  Pressure  To  Discharge  Accumulated  Water  W. 


SEC.  488] 


STEAM  TRAPS 


405 


STEAM  TRAPS  MAY  BE  CLASSIFIED  ACCORDING  To  THE  CHARACTER 
OF  DISCHARGE  as:  (1)  Continuous-discharge  traps,  which  are,  mainly,  of 
the  ball-float  type.  (2)  Intermittent-discharge  traps. 


tout 


Lkfiihl  Filled-*, 
Valve-,  Bourdon  Tubej 


.-Closed  Feeofmter  Heater 

Exhaust  Pipe  from  Engine 

Exhaust  Steam  Met- 


— Regukttiny  Screw 

-—Outlet 


Outlet--' 


FIG.  397. — Marck  Expansion  Steam  Trap  Installed  For  Draining  Closed  Feed  Water 
Heater.  When  Steam  Enters  Casing  Of  Trap,  T  Expands  And  Closes  V.  When  Water 
Accumulates  In  Inlet  Pipe,  T  Contracts  and  Opens  V.  Water  Enters  Casing  C  Of 
Trap  And  Passes  Out  Through  0. 

488.  The  Main  Operating  Principle  Of  Return  Steam- 
Traps  (Fig.  399)  is  equalization  of  pressure  between  the 
interior  of  the  trap  and  the  interior  of  the  boiler  into  which 

?. — -Drains  from  Reheating  Colls  in  Exhaust-  Steam  Receivers-  - . . 
-Drains  from  Separators  in  Steam  Supply  Pipes — ._ 


Seofty- 
. -Valve 
4-  -Plugged  Outlet 

-Valve  Rod-.... 


---Valve  Roof  Housing- - 
Bucket-  • 
Drain  7rap 


Strong 
Trap--- 
Hinge-  -  • 


•Hinge 


FIG.  398. — Arrangement  Of  Tilting-Bucket-Float  Intermittent-Discharge  High- 
Pressure  Steam  Traps  For  Draining  Live-Steam  Separators  And  Reheating-Coils  Of 
Two  Cross-Compound  Engines. 

the  trap  is  intended  to  discharge.  This  is  accomplished  by 
admitting  boiler-steam  to  the  trap.  With  the  equalization 
of  pressure,  the  water  which  has  been  collected  in  the  trap 
flows  out  by  gravity. 


406 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  13 


489.  The  Volume,  In  Cubic  Feet,  Of  Steam  Required  For 
Each  Discharge  Of  A  Return  Trap  is  approximately  equal  to 
the  volume,  in  cubic  feet,  of  the  water  discharged. 


Counterweight  for  Holol/ng 
Bowl  In  Filling  Position 


Live 
Steam 
Pipe 


Hollow  Horn  of  Yote- 

Conveys  Steam 
to  Live  Steam  Pipe 


Pipe  Conveying- 
Condensation- 

Water  to  Trap 


Live  Steam  Vatve, 
Connected  to  Steam 
Space  in  Boiler- •• 


,  Check -Valve, 

'-Opens  Toward 

Trap 


•Check -Valve, 

Opens  from 

Trap 


Pipe  Conveying 

< Water  from  Trap 

to  Wafer  Space 
in  Boiler 


Fia.  399. — Bundy  Return  Trap.  When  B  Fills  With  Water  And  Falls,  V  Opens  And 
A  Closes.  Steam  Then  Passes  Into  Bowl  Through  H  And  L,  And  Water  Is  Forced  Out 
Through  F,  T,  C,  And  D.  When  B  Empties  And  Rises,  V  Closes  And  A  Opens.  Con- 
densation-Water Then  Passes  Into  Bowl  Through  W ,  S,  T,  And  F . 


EXAMPLE. — Assume  that  a  return  trap  is  discharging  into  a  boiler 
under  100  Ib.  pressure.  Then  the  weight  of  the  steam,  which  is  admitted 
to  the  trap  is  (as  taken  from  a  steam  table)  about  0.25  Ib.  per  cu.  ft. 
The  returned  water  of  condensation  weighs  about  60  Ib.  per  cu.  ft.  Now 


SEC.  490] 


STEAM  TRAPS 


407 


as  stated,  above,  1  cu.  ft.  (60  Ib.)  of  water  requires  1  cu.  ft.  (0.25  Ib.) 
of  steam.  Hence,  1  Ib.  of  water  requires  0.25  -r  60  =  0.0042  Ib.  of 
steam. 

Glass    ne~/  ^O»       .•'. -Toggle-Joint  Valye- 

Gage--    zjfa      jZ0*^^./0Pe™t'"yf1echct'1'sm 
Plugged  Or! fice  for  Ac- 
to  Seat  Bushing- 
Blowut 


•Valve-'  Shkld-7^'/. 
r--..    Valve-Seat  .^ 
Orifice--  Bushing..-' *^  Oufle... 


FIG.  400. — American  Ball-Float  Continuous-discharge  High-pressure  Steam  Trap. 

NOTE. — A  portion  of  the  heat  of  the  steam  is  lost  by  radiation  from  the 
trap.  Also,  steam  may  be  lost,  at  each  discharge,  through  the  vent- 
valve  (A,  Fig.  399).  The  cumulative  loss  from  these  sources  may  amount 
-to  1  per  cent,  of  the  total  evaporation  of  the  boiler. 


'•Monel  Metal 

Valve  Stem      Housinct1     L J 

Housing/.      Phosphor-Bronze'. 

Valve-Seat  Bushing/' 

FIG.  401. — Toggle-Joint  Valve-Operating  Mechanism  Of  American  Ball-Float  High- 
Pressure  Steam  Trap. 


490.  The  Economy  Of  Return  Steam-Trap  Service  resides, 
mainly,  in  the  saving  effected  by  returning  the  water  of  con- 
densation from  high-pressure  steam  apparatus  directly  to 
the  boilers,  instead  of  returning  it  thereto  in  relays,  as  through 
a  receiver  or  feed-water  heater  under  atmospheric  pressure. 

EXPLANATION. — In  industrial  processes  which  require  steam  for  heat- 
ing, drying  or  boiling,  the  steam  is  commonly  supplied  from  the  boilers, 
and  is  condensed  in  the  manufacturing  apparatus  under  pressures  rang- 
ing from  a  few  pounds  up  to  100  pounds  or  more. 

Where  steam  of,  say,  80  Ib.  pressure  is  used  in  heating-coils,  as  in  a 
high-temperature  dry-room,  the  water  of  condensation  may  leave  the 


408  STEAM  POWER  PLANT  AUXILIARIES         [Div.  13 

coils  at  a  temperature  of  300  deg.  fahr.  If  such  water  is  trapped  to  an 
open  receiver  or  feed-water  heater,  it  will,  immediately  it  is  discharged 
by  the  trap,  expand  and  cool  to  the  boiling  point  under  atmospheric 
pressure.  Also,  its  temperature  must  be  still  further  reduced  to  about 
210  deg.  fahr.  in  order  that  its  delivery  to  the  boilers,  by  a  feed-pump, 
may  be  facilitated.  Thus  the  water  will  have  thrown  off  the  heat  cor- 
responding to  a  temperature  reduction  of  about:  300  —  210  =  90  deg. 
fahr.  Furthermore,  it  will  have  lost  a  considerable  portion  of  its  own 
bulk  and  some  heat  through  vaporization.  While  most  of  the  water 
thus  vaporized  may  be  recovered,  some  of  it  will  be  a  dead  loss. 

The  saving  that  might  be  realized,  in  this  case,  by  returning  the  water 
of  condensation  directly  from  the  heating-coils  to  the  boilers  is,  there- 
fore, represented  by  (1)  The  quantity  of  coal  required  to  supply  the  heat 
corresponding  io  a  temperature  reduction  of  90  deg.  fahr.  plus  (2)  The  heat 
lost  through  vaporization. 

491.  A  Proper  Location  For  A  Return  Steam  Trap  (Fig.  399) 
is  at  least  3  ft.  above  the  normal  water-level  in  the  boiler  to 
which   the   trap    is   attached.     This   will   insure   a   positive 
gravitational  flow  of  the  returned  water  from  the  trap  to  the 
boiler. 

492.  The   Economy  Of  Non -Return  Steam -Trap  Service 
subsists,  mainly  in  the  saving  effected  by  preventing  the  steam 
from  blowing  through  drips  and  drains  directly  to  the  atmos- 
phere.    It  is  contingent  upon  two  principal  considerations. 
(1)  Selection  of  the  proper  type  of  trap  for  the  particular  service 
requirements.     (2)   The  area  of  the  trap  discharge-valve  orifice 
and  the  condition  of  the  valve. 

EXAMPLE. — Where  a  %-in.  drain  pipe  from  a  steam-piping  system 
under,  say,  165  Ib.  per  sq.  in.  gage  pressure  is  blowing  directly  to  the 
atmosphere,  the  resulting  loss  of  steam  may  amount  to  about  1,120 
pounds  per  hour.  This  is  the  equivalent  of,  approximately,  35  boiler 
horse  power.  Assuming  that  a  boiler  horse  power  costs,  say,  $3.25  per 
mo.,  the  total  monthly  loss  from  this  source  will  amount  to  about  3.25  X 
35  =  $113.75.  With  the  drain-pipe  connected  to  a  properly-selected 
steam  trap,  the  loss  of  steam,  due  to  condensation  in  the  drainage 
connections,  might  be  reduced  to  about  32  pounds  per  hour. 

NOTE. — THE  AREA  OP  THE  VALVE-ORIFICE  OF  A  TRAP  FOR  LOW- 
PRESSURE  SERVICE  should  equal  the  cross-sectional  area  of  the  size  of 
pipe  for  which  the  outlet  orifice  of  the  trap  is  tapped. 

THE  AREA  OF  THE  VALVE-ORIFICE  OF  A  TRAP  FOR  MEDIUM  OR 
ORDINARY  HIGH-PRESSURE  SERVICE,  as  where  the  drainage  from  a  live- 
steam  separator  is  discharged  into  an  open  feed-water  heater,  may  be 


SEC.  493]  STEAM  TRAPS  409 

smaller  than  the  openings  in  the  pipe-connections.  It  should,  however, 
in  any  case,  be  large  enough  to  obviate  liability  of  the  passage  becoming 
clogged  with  particles  of  scale. 

493.  Steam  Traps  Which  Are  Dependent  Upon  Temperature 
Changes  For  Their  Operation  Should  Not  Serve   Separators 
Or  Similar  Apparatus,  in  the  draining  of  which  the  trap  should 
operate  instantly  after  the  accumulated  water  has  attained 
the  head  at  which  it  should  discharge. 

EXPLANATION. — Assume  that  either  a  float-operated  trap  or  a  tilting 
trap  is  installed  for  draining  the  steam-separator  in  a  supply  line  which 
ordinarily  conveys  90-lb. -pressure  steam.  The  trap  will  continue  to 
function,  without  intermission,  if  the  pressure  rises  to,  say,  100  Ib.  or 
falls  below  90  Ib.  But  if  an  expansion  trap  is  substituted,  it  must, 
necessarily,  be  set  to  open  at  the  temperature  of  the  condensation 
from  the  90  Ib. -pressure  steam,  which  may  be  as  low  as  310  deg.  fahr. 
Consequently,  if  the  pressure  rises  to  100  Ib.,  at  which  the  condensation 
may  reach  the  trap  at  about  320  deg.  fahr.,  the  expansion  trap  will 
remain  closed  until  the  temperature  of  the  condensed  water  drops  to 
310  deg.  fahr.  During  the  requisite  time-interval,  however,  the  con- 
densate  accumulation  might  become  dangerously  excessive.  On  the 
other  hand,  if  the  pressure  falls  below  90  Ib.,  condensation  may  reach  the 
trap  at  some  temperature  below  310  deg.  fahr.  Hence,  the  expansion 
trap  will  blow  steam  so  long  as  the  diminished  pressure  continues. 

494.  Steam-Traps    For    Attachment    To    Heating    Coils 

(Fig.  395)  and  similar  apparatus,  may,  with  advantage, 
operate  on  the  principle  of  thermal  expansion. 

EXPLANATION. — Assume  that  steam  at  90  Ib.  pressure  is  circulated  in  a 
set  of  heating  coils.  Then  the  water  of  condensation  will  form  at  a 
temperature  of  about  330  deg.  fahr.  Hence,  a  float-operated  trap  or 
a  tilting  trap  will  discharge  the  water  at  approximately  this  temperature. 
Assuming  that  the  surrounding  air  is  heated  to  150  deg.  fahr.,  the 
quantity  of  heat  in  the  trap-discharged  water  which  will  be  rendered 
unavailable  for  radiation  from  the  coils  will  correspond  to  a  temperature 
range  of  330  —  150  =  180  deg.  fahr.  But  if  an  expansion  trap  is  sub- 
stituted, it  may  be  set  to  discharge  at  150  deg.  fahr.  Thereby  the  maxi- 
mum available  thermal  value  of  the  steam  delivered  to  the  coils  will  be 
realized  for  heating. 

495.  The  Proper  Location  For  An  Ordinary  High-  Or  Low- 
Pressure  Steam  Trap  (Figs.  394  to  398)  is,  with  reference  to 
the  location  of  the  apparatus  which  the  trap  is  intended  to 
serve,  such  that  the  drainage-water  will  flow  to  it  by  gravity. 


410 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  13 


NOTE. — If  the  apparatus  to  be  drained  is  located  at  an  inconveniently- 
low  elevation,  as  on  the  bottom  of  a  narrow  pit  or  trench,  an  expansion 
trap  may  be  located  (Fig.  402)  at  a  higher  elevation  if  the  drainage  water 
leaves  the  apparatus  under  sufficient  pressure.  There  should  be  at  least 
^2  Ib.  per  sq.  in.  pressure  for  each  foot  vertical  height. 

EXAMPLE. — A  steam-pressure  of  5  Ib.  per  sq.  in.  in  the  heating-coil 
(Fig.  402)  will,  practically,  balance  a  column  of  water:  5  -f-  0.5  =  IQft. 
high.  Hence,  the  water  of  condensation  will  be  forced  to  the  trap,  if 
the  trap-inlet  is  located  less  than  about  10  ft.  above  the  drainage-outlet 
of  the  coil. 

Trap  Discharge  Pipe---. 
By-Pass - 


Steam-Trap^ 
<-5te&m-Supply 
P       i        t 


Flo.  402. — Method  Of  Trapping  Condensation  From  Heating  Coil  Located  On  Bottom 

Of  Deep  Pit. 


496.  The  Location  Of  An  Expansion  Trap  should  be  such 
that  its  operation  will  not  be  affected  by  excessive  variations 
of  temperature  occurring  in  the  surrounding  atmosphere. 

497.  The  Capacity  Of  A  Steam-Trap  may  be  rated  (Table 
498)  either  in  terms  of  the  quantity  of  water  to  be  trapped  per 
hour,  or  in  terms  of  the  extent  of  radiating  surface  in  the  ap- 
paratus from  which  the  trap  may  drain  water  of  condensation. 

NOTE. — It  is  commonly  assumed  that  each  square  foot  of  direct  ra- 
diating surface  in  a  heating  system  will,  ordinarily,  condense  about  0.33 
Ib.  of  steam  per  hour.  It  is  also  assumed  that  the  radiation  from  each 
lineal  foot  of  1-inch  pipe  in  a  heating  coil  will,  ordinarily,  condense  about 
0.19  Ib.  of  steam  per  hour.  Where  very  wet  products  are  to  be  dried 
in  a  kiln  or  dry-room,  a  trap  for  draining  the  heating  coils  should  be 
selected  on  a  basis  of  0.56  Ib.  of  steam  condensed  per  hour  per  lineal  foot 
of  1-inch  pipe.  Where  the  heated  air  is  circulated  under  pressure  of  a 
fan-blower,  the  basis  of  selection  should  be  0.94  Ib.  of  condensation  per 
hour  per  lineal  foot  of  1-inch  pipe. 


SEC.  498] 


STEAM  TRAPS 


411 


498.  Table  Showing  Dimensions  And  Capacities  Of  Steam- 
Traps  Working  Under  Medium  Pressure  (Adapted  from 
Swendeman's,  A  STEAM-TRAP  CATECHISM). 


Rated  capacities  per  hour 

Size,  in 

Steam 

Diam.,  in 

in.,  of 

pressures, 

in.,  of  valve 

pipe 

inlb. 

Gal.  of 

Pounds  of 

Lineal  feet 

Sq.  ft.  of 

orifice 

connec- 

per sq.  in. 

water 

water 

of  1-in. 

radiating 

tions 

(Gage) 

dis- 

dis- 

pipe 

surface 

charged 

charged 

drained 

drained 

50 

375 

3114 

5538 

1846 

>4 

H 

75 

459 

3811 

6776 

2258 

100 

530 

4402 

7827 

2609 

125 

593 

4976 

8847 

2949 

50 

584 

4847 

8618 

2873 

.He 

y* 

75 

715 

5936 

10554 

3518 

100 

826 

6853 

12184 

4062 

125 

923 

7662 

13624 

4542 

50 

709 

5883 

10460 

3486 

ly** 

H 

75 

868 

7205 

12810 

4270 

100 

1002 

8320 

14793 

4931 

125 

1122 

9302 

16540 

5514 

50 

844 

6998 

12442 

4147 

H 

1 

75 

1034 

8579 

15754 

5085 

100 

1194 

9986 

17692 

5898 

125 

1334 

11075 

19692 

6564 

50 

1149 

9535 

16954 

.  5651 

K* 

1>£ 

75 

1407 

11680 

20767 

6922 

100 

1625 

13486 

23978 

7993 

125 

1816 

15073 

26799 

8933 

50 

1501 

12537 

22290 

7430 

y* 

tii 

75 

1838 

15252 

27118 

9039 

100 

2122 

17616 

31322 

10441 

125 

2363 

19694 

35017 

11672 

499.  The  Quantity  Of  Condensation-Water  To  Be  Trapped 
From  A  Piping  System  may  be  approximately  computed  by 
the  following  formula : 

(106)  Ww  =  AfK  (Ib.  per  hr.) 

Wherein:  Ww  =  weight  of  condensation  in  pounds  per  hour. 
Af  =  area  of  piping  surface,  in  square  feet.  K  =  conden- 
sation, in  pounds  per  hour  per  square  foot  of  pipe  surface, 


412 


STEAM  POWER  PLANT  AUXILIARIES         [Div.  13 


corresponding  to  the  observed  steam  pressure,  as  given  in 
Table  500. 

500.  Table  Showing  Rate  Of  Condensation,  In  Uncovered 
Pipe  Lines,  Of  Steam  At  Various  Pressures.  Adapted  from 
Elliott  Companys'  BULLETIN  G  ON  STEAM-TRAPS. 


Steam    pressure,     in 

Ib.  per  sq.  in.  (gage) 

5 

10 

20 

30 

40 

50 

60 

80 

100 

125 

Condensation,    in  Ib. 

per  hr.,  per  sq.  ft.  of 

pipe-surface  

0.7 

0.8 

0.9 

1.0 

1.1 

1.2 

1.3 

1.6 

1.7 

1.9 

EXAMPLE. — It  is  found,  by  computation  that  the  high-pressure  piping 
in  a  boiler  and  engine  plant  exposes  2,683  sq.  ft.  of  radiation-area.  The 
steam  pressure  is  115  Ib.  per  sq.  in.,  gage.  What  size  of  trap,  as  listed 
in  Table  498  should  be  used  for  draining  the  system? 

SOLUTION. — By  Table  500,  the  condensation  rate  for  steam  at  100  Ib. 
pressure  =  1.7  Ib.  per  hr.  per  sq.  ft.  of  exposed  surface,  and  for  steam 
at  125  Ib.  pressure  =  1.9  Ib.  per  hr.  per  sq.  ft.  of  exposed  surface.  Hence, 
the  condensation  rate  for  steam  at  115  Ib.  pressure  =  (1.9  —  1.7)  -f- 
(125  -  100)  X  (115  -  100)  +  1.7  =  1.82  Ib.  per  hr.  per  sq.  ft.  of  ex- 
posed surface.  Applying  For.  (106)  Ww  =  A  fK  =  2683  X  1.82  = 
4,883.06  Ib.  per  hr.  Hence,  by  Table  498  a  K-in.  trap  having  a  M-in. 
valve  orifice  should  be  used. 

601.  The  Piping  Of  A  Steam-Trap  should  be  adapted  to  the 
particular  service  for  which  the  trap  is  installed.  Numerous 
right-angled  turns,  and  runs  of  excessive  length  in  the  dis- 
charge piping,  should  be  avoided.  To  obviate  interference, 
the  discharges  from  low-pressure  and  high-pressure  traps 
should  be  piped  independently. 

NOTE. — EVERY  STEAM  TRAP  SHOULD  HAVE  AN  EXTERNAL  BY-PASS 
(B,  Fig.  394).  Also,  stop  valves,  V\  and  F2,  should  be  inserted  between 
the  by-pass  connections  and  the  inlet  and  outlet  orifices  of  the  trap. 

STRAINERS  IN  TRAP-!NLET  CONNECTIONS  (B,  Fig.  395)  may  be  used  to 
prevent  particles  of  scale,  or  other  solid  substance,  from  entering  the 
trap  and  fouling  the  valve. 

PROVISION  FOR  DRAINING  TRAP  DISCHARGE-PIPES,  while  the  traps 
are  inoperative,  (D,  Fig.  398)  should  be  made  when  the  traps  are  ex- 
posed to  freezing  in  cold  weather. 


SEC.  502]  STEAM  TRAPS  413 

502.  Check-Valves  Should  Be  Inserted  In  The  Discharge 
Pipes  Of  Steam  Traps  where  two  or  more  high-pressure  traps 
discharge  (Fig.  398)  into  a  common  discharge-line  or  where 
a  return-trap  (Fig.  399)  is  used  for  boiler-feeding. 

NOTE. — For  ordinary  high-pressure  service,  the  check-valves  (C,  Fig. 
398)  in  the  discharge  pipes  of  steam-traps  may  be  of  standard  weight  and 
may  be  filled  with  renewable  composition  discs.  For  boiler-feed  ser- 
vice however,  the  check-valves  (S  and  Q,  Fig.  399)  should  be  extra 
heavy  and  should  have  solid  brass  discs.  Check-valves  with  composi- 
tion discs  are  ill-adapted  to  withstand  the  stresses  of  boiler-feed-  service. 

503.  A  Vent-Pipe  Connecting  A  High-Pressure  Trap  With 
The  Apparatus  Drained  (P,  Fig.  394)  is  often  necessary  to 
insure  regular  operation  of  the  trap. 

EXPLANATION. — With  a  scant  flow  of  water  from  the  separator  (S, 
Fig.  394)  the  upper  part  of  the  trap  will  contain  steam  of  the  same 
pressure  as  that  in  the  separator.  Should  a  slug  of  water  enter  the  sepa- 
rator, direct  communication  between  the  steam-occupied  space  in  the 
trap  and  the  steam  space  in  the  separator  will,  in  the  absence  of  a  vent 
pipe,  be  closed.  The  flow  from  the  separator  will,  therefore,  cease  until 
the  steam  in  the  trap  condenses.  Restoration  of  an  unimpeded  flow  may 
be  further  delayed  by  air  mingled  with  the  trapped  steam. 

604.  The  Care  of  Steam  Traps  involves  periodic  inspections 
and,  when  necessary,  repair  or  replacement  of  the  valves  or 
seats.  Inspection  should  be  made  frequently  because  the 
flow  of  water  through  steam  traps  cuts  into  the  valves  and 
seats,  which  may  then  leak  or  "blow"  steam.  Since  the 
traps  are  enclosed — as  are  usually  the  discharge  pipes — a  leak 
would  not  ordinarially  be  noticed.  But  by  placing  the  ear  to  a 
trap,  the  blowing  can,  frequently,  be  detected.  A  still-better 
method  for  detecting  the  leaks  consists  of  providing  an  opening 
in  the  discharge  pipe,  from  which  the  leak  is  then  visible. 
Since,  as  stated  in  Sec.  492,  losses  from  leaks  readily  become 
excessive  and  expensive,  a  leaky  trap  should,  immediately, 
be  taken  from  service  and  repaired  upon  discovery  of  the  leak. 

QUESTIONS   ON  DIVISION   13 

1.  What  are  the  general  uses  of  steam  traps? 

2.  What  is  the  distinction  between  a  return  trap  and  a  non-return  trap? 

3.  What  types  of  traps  operate  on  the  principle  of  buoyancy? 

4.  Through  what  media  is  the  expansion  principle  utilized  in  the  operation  of  steam- 
traps?     (See  Fig.  397). 


414  STEAM  POWER  PLANT  AUXILIARIES         [Drv.  13 

5.  What  is  the  essential  operating  principle  of  return  traps? 

6.  What  approximate  volumetric  ratio  exists  between  the  water  discharged  by  a 
return  trap  and  the  steam  required  to  operate  the  trap. 

7.  What  are  the  apparent  sources  of  loss  of  heat  energy  in  the  operation  of  return 
traps? 

8.  How  is  the  economy  of  return-trap  service  principally  manifested? 

9.  What  is  the  minimum  effective  elevation  of  a  return  trap  with  reference  to  the 
boiler  it  is  intended  to  feed? 

10.  What  considerations  mainly  affect  the  economy  of  non-return  trap  service? 

11.  Why  are  expansion  traps  inadaptable  for  draining  live-steam  separators? 

12.  What  types  of  traps  should  be  connected  to  live-steam  separators? 

IS.  Why  are  expansion-traps  well    adapted    for    draining    high    pressure    heating 
apparatus? 

14.  What  is  the  proper  location  for  a  non-return  steam  trap  relative  to  the  elevation 
of  the  apparatus  it  is  intended  to  drain? 

15.  Under    what    conditions    might  an  expansion    steam-trap    be  located    above  the 
apparatus  it  is  intended  to  drain? 

16.  What  are  the  common  bases  of  rating  for  steam-traps? 

17.  Mention  five  structural  features   of  general  importance  in  the  piping  of  steam 
traps. 

18.  Under  what   circumstances  are   check-valves  needed  in  the  discharge  pipes    of 
steam-traps? 

19.  Explain  the  purpose  of  a  vent  pipe  connecting  the  top  of  a  steam  trap  with  the 
top  of  a  steam  separator. 

PROBLEMS   ON  DIVISION   13 

1.  If  a  steam  trap  is  13  ft.  above  the  apparatus  to  be  drained,  what  pressure  will  be 
required  to  force  the  water  up  to  the  trap? 

2.  It  is  found  that  a  certain  uncovered  pipe  line  has  a  surface  of  4530  sq.  ft.     The 
steam  in  the  line  is  at  80  Ib.  per  sq.  in.  gage  pressure.     What  size  trap  should  be  used? 


SOLUTIONS  TO  PROBLEMS  ON  DIVISION  1 
PUMP  CALCULATIONS 

1.  By  Sec.  1,  9  X  22  -5-  14.7  =  13.5  ft. 

2.  By  For.  (1),  P  =  =        j19  =  13  Ib.  per  sq.  in. 


3.  Total  length  of  straight  pipe  =  115  +  38  =  153  ft.     Three  90  deg. 
elbows  =  3  X  8  =  24  ft.  of  pipe.     Two  plugged  tees  =  2  X  16  =  32  ft. 
of  pipe.     Two  globe  valves  =  2  X  8  =  16  ft.  of  pipe.     Total  equivalent 
pipe     length  =  153+24+32  +  16  =225    ft.     Total    friction     head, 
Lh/T  =  (225  -f-  100)  3.70  =  8.32  ft.  head.     Head  equivalent  to  150  Ib.  per 
sq.  in.,  LhmP  =  150  X  2.31  =  346  ft.  head.     Measured  head  due  to  lift, 
Lhmd  =  38  ft.  head.     Total  measured  head,  LhmT  =  346  +  38  =  384 
ft.  head.     Total  head  on  pump,  LhT  =  Lh/T  +  LhmT  —  8.32  +  384  = 
392.32  ft.   head   (neglecting  velocity  head).     By  For.    (1)   P  =  LhT  + 
2.31  =  392.32  +  2.31  =  170  Ib.  per  sq.  in. 

4.  Length  of  straight  pipe  =  153  ft.     Three  90  deg.  elbows  =3X6  = 
IS  ft.  of  pipe.     Two  plugged  tees  =  2  X  12  =  24ft.  of  pipe.     Two  globe 
valves  =  2  X  6  =  12  ft.  of  pipe.     Total  equivalent  length  of  pipe  = 
153  +  18  +  24  +  12  =  207    ft.  =  equivalent     length     of     pipe.     Total 
measured  head  =  LhmT  =  384  ft.     Head  delivered  by  pump  (Prob.  3), 
LhT  =  392  ft.   head.     Head  available  as  friction  head  Lh/T  =  LhT  — 
LhmT  =  392  -  384  =  8  ft.  head.     Friction  head  available  per  100  ft.  of 
pipe  =  8  -r-  (207  -i-  100)  =  3.86  ft.  head.     From  Table  14,  the  water 
delivered  =  about  6^  gal.  per  min.     (This  is  found  by  interpolation.) 

6.  90  cu.  ft.  per  min.  =  90  X  7.48  =  673.2  gal.  per  min.  By  For.  (7), 
di  =  4.95\/Vgm/vm  =  4.95  A/673.  2/2  10  =  8.9  in.  or  a  9-in.  suction  pipe 
would  be  selected  and  4.95  \/673.  2/390  =  6.5  in.,  or  a  7-in.  discharge 
pipe  would  be  selected. 

6.  By  For.  (14),  Vc/  =  LAN,/1,728  =  20  X  102  X  0.7854  X  65  X  2  -i- 
1.728  =  118.2  cu.  ft.  per  min. 

7.  By  For.   (17),  X  =  [100(FC/  -  F.)]/Fe/  =  [100  X  (510  -  487]  -,'- 
510  =  4.51  per  cent. 

8.  By    For.     (18),    Ev  =  100Va/Vc/  =  100  X  487  -=-  510  =  95.5    per 
cent. 

9    Bv  For   (19)    V    -  ^LrE"d  -  3.52  X  (110  X  6.5  4-  12)  X  0.98  _ 
9)'  Va          183.35  "  183.35 

3.9  cu.  ft.  per  min. 


By  Fo,  <20),  4.  - 


3.97,  or  practically  4  in. 
415 


416  SOLUTIONS  TO  PROBLEMS 

11.  By  For.  (22):  vm  =  dpzLt/di2  =  52  X  80  -H  22  =  500  ft.  per  min. 

12.  By  For.    (23),    Wu  =  WLhmT  =  20,106  X  38.5  =  774,081  ft.-lb. 


13.  By  For.  (24),       *  -  -  -  23.5  h.P. 


1d      R      Fn       «m     P  m  9,500    X   310 

14.  By  For,  (30),  Pbhp  =  =  "          =  X°5  h'p' 


<*    r,     IT       ,o^    T.         100WJW       100  X  9,000,000  X  120 
16.  By  For.  (31),  Dc  = ^-  3500 

30,857, 143  ft'.-lb.  per  100  Ib.  of  coal. 


SOLUTIONS  TO  PROBLEMS  ON  DIVISION  2 
DIRECT-ACTING  STEAM  PUMPS 

1.  The   effective  plunger  area  is  (122  X  0.7854)  -  [(32  X  0.7854)  -f- 

2]  =  109.6  sq.  ;n.  The  area  of  opening  of  each  valve  is  0.25  X  4  X 

3.14  =  3.14  sq.  in.  By  Sec.  (60),  109.6  X  0.3  -i-  3.14  =  10.5,  or,  prac- 
tically, 11  valves. 

SOLUTION  TO  PROBLEMS  ON  DIVISION  3 
CRANK-ACTION  PUMPS 

1.  Substituting  directly  in  For.  (48):— 

Pbhp  =  ^M^f  =  15°l,3of5  =  26  h-  P-  (Use  25  *• 

2.  Vgm  (For.  49)  =  30  X  0.9  =  27  gal.  per  min. 

LhmT  =   50  +175  ft.  =  225ft. 
K  (Table  108)  =0.69 

LfK  =  0.69  X  175  =  121  ft. 

Substituting  in  For.  (49),  Pbhp  =  ^^Jp12^  =  4.7  h.  p. 

(Use  5-h.p.  motor). 

SOLUTIONS  TO  PROBLEMS  ON  DIVISION  4 
CENTRIFUGAL  AND  ROTARY  PUMPS 

1.  By    For.     (51),     the    velocity  =  vm  =  481  \/L/  =  481  X  Vl60  = 
6,085  ft.  per  min. 

2.  The    circumference   of   the  impeller  =  6,085  -J-  1710  =  3.558  ft.  or 
3.558  X  12  =  42.7  in.     The  diameter  =  circumference  +  3.1416  =  42.7 
-7-  3.1416  =  13.6  in. 

3.  By    For.     (53),  the    head    produced    at   the    new    speed  =  L/,7-2  = 


SOLUTIONS  TO  PROBLEMS  417 

4.  By.  For.  (52),  the  quantity  of  water  delivered  at  the  new  velocity  = 

N,  X  Vgmi      1,600  X  400 
Vgm2  =          N^  1,450~     =  441    gal.    per    mm. 

6.  By    For.     (61),    the    width    of    a  single  belt  =  Lw  =  ->5       *     b*P 


2,520  X  10 
900X7 


SOLUTIONS  TO  PROBLEMS  ON  DIVISION  6 
INJECTORS 


From  For.  (62) 
1    w      -  x 

1  .      W  a  ,w    — 


Here  T/<    =  60  deg.   fahr.    and  Tfd    =    200  <%.  /a&r.   For   100  Ib. 
per  sq.  in.  gage  the  following  valves  are  found  in  the  steam  tables: 
Tfa  =  338  deg.  fahr.,  Hv  =  879.9  B.t.u.  per  Ib.     When  there  is  a  moisture 
content  of  2>£  per  cent.,  x  =  1.00  —  0.025  =  0.975.     Then  substituting 
in  For.  (62)  : 

0.975  X  879.9  +  (338  -  200) 

Wsw  =  —  —  ^7^  —  —^  ---  •••7.11  Ib.  of  water  pumped  per  Ib. 

200  —  ou 

of  steam 
Again  applying  For.  (62)  :  — 

o    in  -  0.975  X  879.9  +  (338  -  Tfd) 
!T,d  -60 

Transposing  and  simplifying: 

10  Tfd-  600  =  857.9  +  338  -Tfd 

11  17/*-  1,796 
Therefore: 

Tfd  =  163.26  deg.  fahr. 

3.  From  For.  (69),  for  water  tube  boiler: 

Gallons  per  hour  of  injector  =  =  =  206.7 


Increasing  by  30  per  cent,  these  results:  206.7  -f  (30  X  206.7)  = 
268.71  gal.  per  hr. 

Looking  in  Table  194,  the  SIZE  B  is  required  to  pump  260  gal.  per  hr. 

Therefore  it  is  the  size  to  use.  Note.  If  the  lift  is  very  great  (over  15 
feet)  it  is  advisable  to  select  the  next  larger  size  of  injector. 

4.  From  Table  194,  under  PIPE  CONNECTIONS  the  size  given  is  %  in.  for 
the  injector  SIZE  B  of  Prob.  3.  This  is  the  correct  size  for  all  steam  and 
delivery  lines,  except  when  the  run  is  unusually  long.  The  suction  line 
will  be  %  in.  for  an  8  ft.  lift.  For  a  15  foot  lift,  a  1-in.  suction  line  would 
be  recommended.  For  a  lift  of  20  ft.,  it  would  be  advisable  to  use  1^-in. 
pipe  for  the  suction  line. 

27 


418  SOLUTIONS  TO  PROBLEMS 

SOLUTION  TO  PROBLEMS  ON  DIVISION  6 
BOILER  FEEDING  APPARATUS 

1.  By  For.  (74)  gal.  per  hr.  required  =  6  X  PnhP  =  6  X  600  =  3,600 
gal.  per  hr. 

To  retain  the  same  per  cent,  excess  capacity  when  boilers  are  forced 
225  per  cent,  the  capacity  is: 

3,600  X  2.25   =  8,100  gal.  per  hr. 

2.  Pounds  of  water  per  hour  required  by  main  engine  = 

500 


The  auxiliaries  require  10  per  cent,  of  this  or  660. 

The  total  normal  requirement  is  then  6,600  +  660  =  7,260  Ib.  per  hr. 
A  50  per  cent,  excess  over  this  capacity  =  1.5  X  7,260  =  10,890  to.  per 
hr. 

There  are  about  8.34  Ibs.  of  water  in  a  gallon. 

Therefore  the  pump  capacity  in  gallons  =     Q'    .     =  1,305  gal.  perhr. 

o.  o4 


SOLUTIONS  OF  PROBLEMS  ON  DIVISION  7 
FEED-  WATER  HEATERS 

1.  The  quantity  of  exhaust  steam  used  in  heating  the  feed-water  = 
(500  X  20  X  11)  -5-  100  =  1,200  Ib.  per  hr.  The  total  heat  in  the  steam, 
above  32  deg.  fahr.,  is  about  1,150  B.t.u.  per  Ib.  Hence,  by  For.  (77): 
Tfl  Wf  +  0.9WS(#  +32) 


r,«  = 

yy  /  ~T  u.»   yy  « 

Q  v  i  9nn  v  n  u=;n  _i_  a9M 

=  196.4  deg.  Mr. 


0.9  W8 

(90  X  10,000)  4-  [0.9  X  1,200  X  (1,150 
10,000  +  (0.9  X  1,200) 

2.  As  given  in  a  table  of  the  properties  of  saturated  steam,  the  total 
heat,  above  32  deg.  fahr.,  in  steam  at  150  Ib.  persq.  in.,  gage,  is  1,195  B.t.u. 

per  Ib.     Hence,   by  For.    (76),   the  saving  =  Hf  =  ^  —  f*~     7*        100 

n   —  \  L  fi  —  6Z) 

-  1,195  -  (OT-  82)  X10°-"-88  '"«"'• 

3.  As  given  in  a  table  of  the  properties  of  saturated  steam,  the  total 
heat,  above  32  deg.  fahr.,  in  steam  at  125  Ib.  per  sq.  in.,  gage,  is  1,192  B.t.u. 
per   Ib.     Hence,    by    For.    (76)    the   probable   thermal   sairing  =  H/  = 

m         _    m  rt-j  f\    _    -I  cA 

10°  -  X  10°   =  5'8  ^  ^     The 


H  -(Tfl  -  32)  1.192  -(150  -32) 

present    annual    cost    of    the    fuel    supply  =  3.5  X  5  X  300  =  $5,250. 

Hence    the    probable    annual    saving  =  -  —  ^^  ---  =  $304.50.     The 

300  ^  6 
interest    on  the  investment  =  —    x  —  =  $18.     The     annual     cost    of 


SOLUTIONS  TO  PROBLEMS  419 

depreciation  =  --  foo~~""    =  $*8.00.     The  annual  cost  of  maintenance 

and    operation  =  12  X  4  =  $48.     Hence,     the     probable     annual     net 
saving  =  304.50  -  (18  +  18  +  48)  =  $220.50. 

4.  By  Table  278,  U  =  350.    Hence,  by  For.  (81)  A/ 
15,000  X  (200  -  70) 

~~l  —    —  70  =  65'6  Sq-  ft' 

=  350  x  (220- 


6.  By  For.  (78)  the  weight  of  steam  condensed, 


Ws  =  0.9(H  +  32)  -  T/,  +  0.177/2 

(205  -  40)15,000 _ 

0.9(1,150.4  +  32)  -  40  +  (0.1  X  205)  " 

SOLUTIONS  OF  PROBLEMS  ON  DIVISION  8 
FUEL  ECONOMIZERS 

1.  By   Fig.    (266),   the   weight   of   combustion   gases   which   contain 
12  per  cent,  of  CO2  =  13  Ib.  per  Ib.  of  coal.     Also,  the  weight  of  combus- 
tion gases  which  contain  8  per  cent,  of  CO2  =  20  Ib.  per  Ib.  of  coal. 
Hence,  the  leakage-air  =  20  -  13  =  7  Ib.  per  Ib.  of  coal.     The  heat, 
above   the  outside  air-temperature,   which  is  contained  in  the   gases 
leaving  the  boiler  =  (550  -  50)  X  13  X  0.24  =  1560  B.t.u.  per  Ib.  of 
coal.     The  heat,  above  the  outside  air-temperature,  which  is  contained  in 
the  leakage  air  =  (250  -  50)  X  7  X  0.24  =  336  B.t.u.   per  Ib.  of  coal. 
Hence,  the  per  cent,  of  loss  =  336  -f-  1,560  =  21.6  per  cent. 

2.  By  For.  (82)  the  requisite  ratio  =  X  =  Tfo  +  Tfw  =  CwWw/C0Wa 
=  (1  X  8)  -T-  (0.24  X  15)  =  2.22. 

3.  By   Fig.    276,   the  lowest  temperature-difference   consistent  with 
profitable   operation  =  100   deg.    fahr.     Hence    (Sec.    305)    the   lowest 
permissible  temperature  of  the  gases  leaving  the  boiler  =  100  +  377.5  = 
477.5  deg.  fahr.     By  Fig.  277,  the  corresponding  boiler  heating-surface 
=  9  to  10  sq.  ft.  per  h.p. 

4.  By    Fig.    278,    the    least    temperature-difference,    consistent    with 
economy,  between  the  water  and  the  gases  =  40  deg.  fahr.     Hence,  the 
lowest  permissible  temperature  of  the  gases  at  exit  =  200  +  40  =  240 
deg.  fahr. 

5.  The   heat-transfer  =  5.5  X  300  =  1,650    B.t.u.    per   sq.  ft.   per  hr. 
Hence,    the   requisite  area  of  heating-surface  =  50,000  X  50  •*•  1,650  = 
1,515  sq.  ft. 

6.  By   For.    (83),  X  =  100  (T"fw  -  T'fw}/(H  +  32) -!/",„  =  100  X 
(250  -  110)  -T-  (1197.3  +  32  -  100)  =  12.4  per  cent. 

7.  The  annual  cost  of  fuel  without  the  economizer  =  (2;400  X  24  X 
4.3  X  300  -T-  2,000)  X  4.25  =  $157,896.     The    saving   effected    by  the 


420  SOLUTIONS  TO  PROBLEMS 

economizer  =  157,896  X  12.3  4-  100  =  $19421.21.  The  annual  cost  of 
operation,  maintenance  and  depreciation  of  the  economizer  =  12,000  X 
0.15  =  $1,800.  Hence,  the  net  annual  saving  =  19421.21  -  1,800  = 
$17,621.21. 

SOLUTIONS  TO  PROBLEMS  ON  DIVISION  9 
STEAM  CONDENSERS 

1.  By  For.  (84),  the  greatest  possible  thermal  efficiency  non-condensing  = 
_         Tj  -  T2        (450  +  460)  -  (255  +  460)       125 

~TT~  450  +  460  =  9iO  =  24'7  Per 

Also  by  For.  (84),  the  greatest  possible  thermal  efficiency  condensing  = 
Tt  -T2        (450  +  460)  -  (80  +  460)      370 
TET  450  +  460  =910  = 

2.  By  For.   (85)  the  saving  in  power  due  to  condensing  operation  = 

49Phmv      49  X  26.5 


p 


=  16.6  per  cent. 


3.  By  the  graph,  Fig.  286,  the  ideal  steam  consumption  of  the  turbine 
is  14  Ib.  per  h.p.  hr.  non-condensing  and  7  Ib.  per  h.p.  hr.  condensing. 
The  actual  steam  consumption  condensing,  then  = 

^  X  22  =  11  Ib.  per  h.p.  hr. 

4.  The  absolute  condenser  pressure  =  29.8  —  27  =  2.8  in.  of  mercury. 
By  For.  (87),  the  absolute  pressure  = 

D  Phmb  —  Phmv         29.8    —  27  1    no   7L 

Pa  =      -£63~          "2:0-3-     =  °8  lb'  Ver  ^  m' 
The  per  cent,  of  the  possible  vacuum  = 

X  100  =  90.6  per  cent. 


5.  By  For.  (88),  the  volume  of  the  condenser  =  V  =  0.001,43  W8  + 
8.25  =  (0.001,43  X  10,000)  +  8.25  =  22.55  cu.  ft.  One  hour  =  3,600 
sec.  One  cubic  foot  of  water  weighs  62.5  lb.  Therefore,  the  volume  of 
the  cooling  water  = 

10,000  X  36 


3,600  X  62.5 


*  «*•/*•  I*"*- 


10  000 
The  volume  of  the  condensate  =  3  eoo'x  625  =  °'044  cu'^'  per  sec' 


Hence,    the    tail-pipe  area  =  L6         -         =  Q  329    sg>    JL  =  0.329  X 


144  =  47.4  sq.  in.     Therefore,  the  tail-pipe  diameter  =       ~K  or  7.8  in. 
or  approximately,  8  in. 


SOLUTIONS  TO  PROBLEMS  421 

6.  By   Table   345,   the  steam  temperature  corresponding  to   a  27  in. 
vacuum  =  115.06  deg.  fahr.     Hence,  the  temperature  difference  between 
the  discharge  and  the  entering  steam  =  115.06  —  105  =  10.06  deg.  fahr. 

By  Table  JH5,  total  heat  in  steam  at  a  27  in.  vacuum  =  1110.2  B.t.u.  per 
Ib.     By  For.  (89),  the  weight  of  cooling  water  required  = 

Ww=  Ws  H  ~  Tf*  +  32  =  10,000  1110'12Q^"_!°850+  32  =  414,880  Ib.  per 

hr.     One  gallon  =  8.3  Ib.  of  water.     Hence  the  volume  of  cooling-water 

414,880 

required  =  an  vx  0  0  =  8ocJ  gal.  per  mm. 
ou  x  o.o 

7.  As  referred  to  a  30-in.  barometer  (Table  345)  the  degree  of  vacuum  = 
30  —  29.5  +  28  =  28.5  in.  of  mercury.     By  Table  345,  the  total  heat  of 
steam  in.  a  28.5  in  vacuum  =  1,100  B.t.u.  per  Ib.     By  For.  (89),  the  weight 
of  water  required, 

W«  =  WsH y    T^  £  32  =  10,000 1?1  87"^  67+ —  =  523,500  Ib.  per  hr. 

8.  By  For.   (90),  the  quantity  of  heat  to  be  abstracted  from  the  steam, 
Ht  =  W8(H  -  Tfc  +  32)  =  150,000(1095.6  -  80  +  32)  =  157,000,000 
b.t.u.  per  hr.     By  For.  (91)  the  tube  surface  required, 


U,  by  Table  350  =  600.     Tfs,  by  Table  345,  =  82  deg.  fahr.     Hence, 

157,000,000  _ 

Af  "600(82  -MI60  +  77])  - 

SOLUTIONS  TO  PROBLEMS  ON  DIVISION  10 
METHODS  OF  RECOOLING  CONDENSING  WATER 

1.  By  Table  393,  the  relative  humidities  of  the  air  at  entrance  and 
exit  are,  respectively,  55  per  cent,  and  92  per  cent.     By  Fig.  315,  the 
weight  of  saturated  water  vapor  per  cubic  foot  of  air  =  0.001,1  Ib.  at 
70  deg.  fahr.  and  0.002,2  Ib.  at  90  deg.  fahr.     Therefore,  the  moisture 

content  of  the  air  at  entrance  =  -  — ^ —  -  =  0.000,605  Ib.  per  cu.  ft., 

0  002  2  V  Q2 
and  of  the  air  at  exit  =  -       JOQ       -  =  0.002,024  Ib.  per  cu.  ft.     Hence, 

the  quantity  of  water  absorbed  per  cubic  foot  of  air  =  0.002,024  —  0.000,605 
=  0.001,419  Ib. 

2.  By  Sec.  402,  the  area  required  for  a  simple  cooling  pond  =  1,000  X 
120  =  120,000  sq.  ft. 

.       1,000  X  15  X  40 
By  Sec.  411,  the  area  required  for  a  spray  pond  =  •       — ^nn — 

3,000  sq.  ft. 

3.  By  solution  of  Problem  1,  the  quantity  of  water  evaporated,  per  cubic 
foot  of  air-flow  through  the  tower,  =  0.001,419  Ib.     By  Sec.  399,  the  heat 
abstracted,  per  pound  of  water  evaporated,  =  1,000  B.t.u.     1  gal.  =  8.3  Ib. 


422  SOLUTIONS  TO  PROBLEMS 

800  X  8.3  X  20  X  80 
Therefore,  the  volume  of  air-flow  per  minute  =  o^oT,419Xl,OOOxTOO 

74,870.  cu.  ft. 


By  For.  (92),  E  =  100        -  -  100  X  5  ~  44'4  *» 

The  water  lost  by  evaporation  =  74,870  X  0.001,419  =  106.24  Ib.  per 

min.  or  or>n    •       ~  X  100  =  1.6  per  cent. 
5UU  X  o.o 

4.  By  solution  of  Problem  3,  the  volume  of  air-flow  =  74,870  cu.  ft. 
per  min.  By  Sec.  423,  the  allowable  velocity  of  air-flow  =  700  ft.  per  min. 

Therefore,  the  free  area  =       '         =  107  sq.  ft. 

By  Sec.  423,  the  free  area  =  64  per  cent,  of  the  total  horizontal  cross- 

sectional  area.     Therefore,   each  side  of  the  base  =  \/ 

\ 

13  ft. 

6.  By  Table  408,  the  discharge,  per  nozzle,  =  60  gal.  per  min.  or  60  X 
60  X  24  =  86,400  gal.  per  day.  Hence,  the  requisite  number  of  nozzles  = 
40,000.000  _ 

86,400 

By  Sec.  411,  1  sq.  ft.  of  pond  area  will  suffice  to  cool  250  Ib.  of  water  per 

40,000,000  X  8.3 


hour.     1  gal.  =  8.3  Ib.     Hence,  the  requisite  area 


250  X  24 


55,333  sq.  ft. 

SOLUTIONS  TO  PROBLEMS  ON  DIVISION  11 
STEAM-PIPING  OF  POWER-PLANTS 

1.  A  table  of  the  properties  of  saturated  steam  shows  the  density  at 
150  Ib.  pressure,  gage,  to  be  0.363  Ib.  per  cu.  ft.     By  For.  (97)  di  = 


=  5'6  in'  0^  practically  6  in. 


2.  By    For.    (99),    dim  =  Vd*2  +  d<22  +  dtf  +  etc.  = 

V2.52  +  42  +  52  +  72  =  9.8-in.,  or,  practically,  IQ-in. 

3.  By  For.  (100),  Lv  =  114  d>  +  (l  +  ^)  =  114  X  6  -=-  (l  +  ^ 

427.5  in.  or  427.5  -h  12  =  35.6ft.     By  For.  (101)  Le  =  76dt-  -=-  (l  + 

-  76  X  6  4-   (l  +  ~\  =  285  in.  or  285  -4-  12  =  23.75  ft.     Hence,  the 

total   equivalent   pipe-length  =  35.6  X  2  +  23.75  =  94.95  ft. 

4.  A  table  of  the  properties  of  saturated  steam  gives  the  temperature 
at  135  Ib.  pressure,  gage,  as  358.5  deg.  fahr.     A  manufacturer's  table  of 
pipe  sizes  (Nat.  Tube  Co.)  gives  the  outside  diam.  of  an  8-in.  pipe  as 
8.625-in.     By     For.     (103),    Lh  =  0.043 \/dpLpTf  =  0.043  X 

V8.625  X  150  X  (358.5  -  60)   =  26.72ft. 


SOLUTIONS  TO  PROBLEMS  423 

6.  A  table  of  the  properties  of  saturated  steam  gives  the  temperature 
and  latent  heat  oi  steam  at  125  Ib.  pressure,  gage,  as  353.1  deg.  fahr.  and 
867.6  B.t.u.,  respectively.  By  formula  (104)  Wc  =  2.7A(T/S  -  Tfa)  •*• 
Hv  =  2.7  X  2.816  X  30  X  (353.1  -  90)  -h  867.6  =  69.17  Ib.  per.  hr. 

SOLUTIONS  TO  PROBLEMS  ON  DIVISION  12 
LIVE-STEAM  AND  EXHAUST-STEAM  SEPARATORS 

1.  Reduction    of    moisture    content:  98  —  95  =  3    per    cent.     From 
Sec.  457  a  full  saving  of  1  per  cent,  results  for  each  per  cent,  increase  of 
dryness   of   the   steam.     Cost   of   coal   per   year:  30  X  365  X  4.00  = 
$43,800.     Saving  due  to  better  separation:  0.03  X  43,800  =  $1314.00. 

2.  By    For.  (105):  E  =  WOWw/Wt  =  100  X  285  -r-  [5600  X  (1.00  - 
0.93)]  =  72.7  per  cent. 

SOLUTIONS  TO  PROBLEMS  OF  DIVISION  13 
STEAM  TRAPS 

1.  From  Sec.  495,  at  least,  13  X  0.5  =  6.5  Ib.  per  sq.  in. 

2.  By   For.    (106),   Ww  =  AfK  =  4;530  X  1.6  =  7,248  Ib.   per  hr.  = 
water  condensed.     From  Table  498,  a  trap  with  %  in.  pipe  connection 
and  1>i2  in.  valve  orifice  is  required. 


INDEX 


PAGE 


Absorption  in  steam  separators 395 

Absorption  separator 389 

operating  principle 395 

Adjustment  of  injector  water  supply  170 
Air  and   water  in  recooling  system, 
devices    for    bringing    into 

contact 337 

and    water,    power    to    remove 

from  condenser 284 

atmospheric,   relative  humidity, 

how  determined 332 

below    suction-valves,    effect    on 

pump 66 

between  pump  suction  and  dis- 
charge decks  effect 67 

in  suction-line,  accumulation  ...       8 
leakage    in    condenser,    preven- 
tion      312 

leakage  into  economizer  setting  260 
temperature,  effect  on  recooling  345 
unsaturated,    water- vapor    pres- 
sure, now  determined 337 

Air-chamber    charging-apparatus  for 

high   pressure  pumps 51 

connections 49 

for  duplex-pump 50 

for  simplex-pump 50 

function  of 50 

in  pump,  apparatus  for  replen- 
ishing, illustration 49 

in  suction  line,  centrifugal  pump  145 

pump,  water-level  in 50 

steam-pump     with,     illustration     48 
volume,    ratio    to    water-piston 

displacement 51 

Air-cooling  in  surface  condenser 290 

Air-pocket  in  centrifugal  pump  suc- 
tion line .  143 

Air-pump,  condenser,  function 283 

Radojet 308 

simplex-pumps  as 65 

Wheeler- Edwards 309 

Alberger  barometric  condenser  dry 
air-pump  connection  illus- 

trati9n 309 

Alberger  hurling-water  pump,  prin- 
ciple of,  illustration 306 

Alberger      Pump      and      Condenser 
Company's      Catalog,      hot 
water  pump  suct'on  graph  137 
Alberger      Pump      and      Condenser 

Company,  centrifugal  pump  111 
Altitude,  effect  on  pump  suction  ....        1 
Altitudes,    pump  suction-lift  at  vari- 
ous, table 2 

American      Correspondence      School 

on  feed-water  heaters 214 

"American  Society  Mechanical  Engi- 
neers,        Proceedings," 
cooling  pond-area  data  .   342 
"Transactions,"  pipe-sizes 

graph 374 

American  Society  for  Testing  Mate- 
rials, standard  terminology 
for  pipes 367 


PAGE 
Ammonia       condenser,      condensing 

water  for 329 

Anchors  for  steam  piping 380 

Area,  effective,  of  piston 20 

of  plunger . 20 

of  steam  cylinder  to  balance 
given  water  pressure,  for- 
mula    37 

of  water  cylinder  to  balance 
given  steam  pressure,  for- 
mula    37 

Asbestos    heat-insulating    for    closed 

heater 246 

Atmospheric     cooling     device,     effi- 
ciency, formula 334 

ideal  and  actual  range 333 

Austin       baffle-plate       angle       live- 
steam  separator  illustration  391 
baffte-plate  underslot  horiz9ntal 
live-steam    separator,    illus- 
tration      392 

exhaust-steam       separator       for 

vacuum  service,  illustration  392 
reverse-current  live-steam  sepa- 
rator, illustration 390 

Automatic  return  apparatus,  illustra- 
tion     202 

Auxiliaries,     non-condensing    steam- 
driven,  in  condensing  steam 

power  plants 180 

supply    of   exhaust   steam   from  231 


Back  pressure  decreased  by  con- 
denser    282 

Badenhausen     integral      economizer 

or  preheater 255 

Badger    and    Sons    Company,    spray 

nozzle 344 

Baffle-plate         separator,     operating 

principle 392 

Balancing  double-suction  pumps  .    ..    117 

Balancing      piston,      automatic     hy- 
draulic      116 

impeller 115 

Ball-float  steam  trap 407 

Baragwanath  single  flow  condenser . .   294 

Barometer,  effect  on  condenser 285 

Barometric  jet  condenser  adjust- 
ment and  care 315 

Baum  baffle-plate  horizontal- 
steam  separator,  illustra- 
tion    392 

exhaust-steam       separator       for 

vacuum  service,  illustration  392 

Bearings,     centrifugal     pump,     care 

of 151 

Bed-plate      of      centrifugal      pump, 

grouting 141 

Belt  drive,  centrifugal  pumps. 131 

required  width,  centrifugal  pump  131' 

Bend,  long-radius    in    pump  piping, 

illustration 6 

to  take  up  expansion  in  pipe  of 
given  length,  pipe  required 
for  formula 378 


425 


426 


INDEX 


PAGE 
Bends     for     taking     up     expansion 

stresses  in  piping  systems  .  .   370 
Blake-Knowles      closed      feed-water 

heater 210 

open  feed-water  heater 233 

Blower,  soot,  for  economizers 258 

inspection  in  economizer 274 

Blow-off  pipe  for  closed  heater 247 

valve  for  heater 246 

valve,  inspection  in  economizer  274 
Body,  freely  falling,  velocity  formulas  i05 
Boiler,  condensation  returned  to  with 

Holly  loop 383 

connection    to     duplicate     main 

headers 372 

evaporative    capacity    increased 

by  hot  feed-water 212 

feed,    belt-driven    double-acting 

pump  for 78 

centrifugal     pumps     table     of 

capacities 194 

per    cent,    returned    to,    from 

surface  condenser 319 

pump  capacities,  table 193 

pump    governors     on    direct- 
acting     steam     pumps  198-201 

pump  sizes 193 

pumps  for 135 

service,     selection     of    direct- 
acting  steam  pump  for 65 

vertical  duplex-pump,  illustra- 
tion        55 

feed-water,   heat  saved  by  pre- 
heating      212 

preheating        with        exhaust 
steam,      monetary      saving 

resulting 213 

feeding  apparatus 171-205 

devices,     relative     economies, 

table 174 

devices,  relative  efficiencies. ..    173 

centrifugal  pumps 136 

pumps     compared     to     steam 

traps  for 205 

table  of  pump  applications  ...    190 

where  exhaust  steam  is  needed  180 

flexibility  due  to  economizer ....   272 

flue,  water  for,  formula 165 

forced,  life  of 197 

forcing,  feed-pump  capacity  ....    197 

forcing  in  large  plants 197 

heating-surface,  additional,  com- 
pared to  economizer 267 

temperature  difference  f or . .  .   268 
high  stresses  caused  by  cold  feed- 
water  207 

plant,         steam     supplied      by, 

methods  of  distributing.  .  .  .   371 

pressure  for  economizer 273 

pressure-head  in 8 

rating     determining     feed-water 

requirements  for,  formula  . .    195 
strain  due  to  cold  feed- water  ...   160 

tubular,  water  for,  formula 165 

water-tube,  gas  temperature  in, 

llustration 267 

water  tube,  water  for,  formula...    165 

with  hot-water  returns,  feeding  202 

Boiler-efficiency  due  to  economizer  . .    272 

Boiler-feeder,  duplex 204 

Boiler-horsepower,    ratio    to  econo- 
mizer heating-surface 269 

Bourdon  tube  vacuum  gage 284 

Brackets,  wall-,  for  steam  piping ....    380 
Briggs  formulas  for  pressure-drop  in 

steam  pipes 377 

Bucket-float  steam  trap 405 


PAGE 

Buckley  siphon   condenser,   diagram  292 
Bundy     gridiron    horizontal    steam 

separator,  illustration 394 

return  trap,  illustration 406 

trap,  illustration 203 

Buoyancy  steam  trap 403-407 

Burg,    F.    A.     Application    of    Con- 
densers     316 

Burhorn       metallic     cooling      tower 
for     double-pipe     ammonia 

condenser 330 

open  atmospheric  cooling  tower  352 
sheet  metal  cooling  tower  louvres  354 
Burnham   compound   simplex-pump, 

illustration 64 

Butt- weld,  strength  of 366 

Butt-welded  pipes 365 

By-pass       economizer      installation, 

illustration 261 

external,  in  steam-trap 412 

on  boiler-feed  pump,  illustration  186 


Carbon  dioxide  drop  through  econo- 
mizer      260 

Carpenter,    H.    V.     Steam-pipe    size 

chart 373 

Center-head,  marking  striking  point     62 
Centrifugal  and  rotary  pumps.  .   101-153 

force,  illustration 102 

pump,  advantages  of 134 

air-chamber  in  suction  line . .  .    145 

application 95,  135 

at    wrong    capacity    or    head  135 

bed-plate  grouting 141 

belt  drive 131 

boiler  feeding 136 

boiler-feed,  table  of  capacities  194 
capacity    and    impeller    speed 

formula 107 

capacity  heads  and  speeds  126-128 

casing,  draining 151 

change     of     operating     condi- 
tions, effect  of 128 

characteristic  graphs 124 

characteristics       at      desired 

speed,  graph 129 

classification 109 

commercial 103 

comparative  efficiency 112 

compared      to      reciprocating 

pumps 134 

connection  of  suction  nozzle.  .    142 

development  of 101 

direct-connected  motor-driven  131 

direction  of  rotation 139 

disadvantages  of 134 

discharge  pipe,  how  valved. .  .    147 

discharge  piping 145 

drives  for 133 

effect    of    various    factors    on 

efficiency  . 105 

efficiency 126-128 

electric  motor  drive 132 

electrically-driven,  steady  vol- 
tage    151 

explanation 104 

Fairbanks-Morse    multi-stage, 

illustration 113 

flexible  coupling  for  connecting  133 
for  boiler-feed,  selection  of  ...   136 
for      boiler      feeding,      disad- 
vantage     193 

for  condensate  pump,  advan- 
tages     308 


INDEX 


427 


PAGE 

Centrifugal  pump  for  condenser  cir- 
culating pump,  advantages  308 

for  condensing  water 135 

foundation  bolts  for 140 

foundation  for 140 

guide  bearing,  illustration.  .  .  .    122 

handling  hot  water 137 

head,    volume,    R.P.M.    rela- 
tions, graph ? 126 

heads  and  speeds 126-128 

illustration  of  principle 103 

impeller,  theoretical  speed  ....  106 
in  connection  with  feed-water 

heater 136 

independent  suction  lines  for, 

illustration 143 

installation 138 

left-hand,    illustration,    defini- 
tion      140 

leveling 141 

maintenance 151 

mechanical  efficiency 193 

methods  of  driving 131 

methods  of  priming 147 

motor-driven,    for   boiler   feed   192 

speed  variation 151 

speeds 132 

multi-stage,  for  high  heads  ...    112 
open-type    impeller,     illustra- 
tion        118 

operating  principle 101 

performance  characteristics.  .  .  122 
positions  of  discharge-nozzles.  140 
power  required  and  speed  of 

impeller,  formula 108 

power  required  to  drive  at  any 

speed 130 

pressure    head    and    speed    of 

impeller 108 

priming 146 

put      in       service,       bearings 

cleaned 151 

quantity    of    water    delivered  107 

required  belt  width 131 

right-hand,  illustration,  defini- 
tion      139 

run  in  wrong  direction 151 

with  discharge  valve  closed  149 

with  casing  empty 147 

selection  of 1 38 

single-  and  double-suction.  ...    114 
single-stage     volute,     illustra- 
tion      103 

size  and  capacity 107 

speeds  heads  and  capacity.  126-128 

starting .  .  .  ; 149,  151 

submerged  type 120 

submersible  type,  illustration.  150 
suction  lift  of,  how  measured  .  8 

piping 141 

vacuum  pipe 145 

test 122-124 

test  conditions  for 124 

theory 101 

two-stage  double-suction  tur- 
bine, illustration 110 

explanation 114 

vertical  shaft 120 

submerged,      thrust-bearing 

for 152 

thrust  bearing  for 121 

volute    and    turbine,    applica- 
tions of Ill 

water  rate  of  turbine  for 193 

with    branch    discharge    pipe, 

illustration 149 

separator 389 


PAGE 

Centrifugal  separator,  operating  prin- 
ciple .      390 

Chamber,  vacuum,  in  pump,  func- 
tion of 51 

Check  valve,  faulty  boiler 169 

in      centrifugal      pump      dis- 
charge     146 

in    discharge    pipes    of    steam 

traps 413 

Checkerwork,  cooling  tower,  com- 
puting height 360 

cypress  board,  illustration 351 

for  mixing  chamber 352 

wood,  for  cooling  towers 353 

Circulating  pump,  surface  condenser  321 
Cleaning      condenser,      cost,      what 

determines 324 

jet      and      surface      condensers, 

relative  cost 323 

Coal    consumption,    duty    of    steam 

pump  on  basis  of,  formula     32 

with  economizer,  graphs 271 

"Coal  Miner's  Pocketbook"  on  pump 

management 66 

Cochrane  feed-water  heater,  illustra- 
tion    226 

heater     in      condensing      plant, 

illustration 180 

horizontal      receiver  -  separator, 

illustration 387 

open        induction        feed-water 

heater 219 

Coefficient  of  heat  transference, 
value  effected  by  conditions 

in  condenser  tubes 304 

Coke  heater  packing 246 

Column,  fluid,  static  head  definition       3 
of     water,     converting     to     unit 

pressure,  formula 4 

Combining  tube  of  injector 156 

Combustion-gas,  temperature,  loss 
of,  ratio  to  gain  of  feed- 
water  temperature 266 

Commonwealth     Edison     Company, 
Fisk    Street   Station,    boiler 
and    economizer    surfaces      269 
Condensate  from  heating,   traps  for 

returning 203 

issuing      from      jet      condenser, 

velocity 299 

Condensation     and     entrainage     in 

steam-piping,  steam  trap  for  403 
due   to   loss   of   heat   from    bare 

steam  pipes  formula 381 

from    heating    coil,     method    of 

trapping,  illustration 410 

in     high-pressure     steam-piping 
returned     to     boilers     with 

Holly  loop. 383 

in  primitive  steam  engines 278 

in     steam-heating     and    power 

apparatus,  steam  trap  for  . .    403 
rate   of  steam   at   various   pres- 
sures    in     uncovered     pipe 

lines,  table 412 

Condensation-water,  quantity  to  be 
trapped  from  piping  system, 

formula 411 

Condenser,  Alberger  barometric,  dry 
air  pump  connection,  illus- 
tration    309 

ammonia,  condensing  water  for  .  329 
low-temperature  cooling  tower 

test  data,  table 358 

temperature  of  water  leaving  332 
and    spray-cooling    outfit,     per- 
formance guarantees 347 


428 


INDEX 


PAGE 

Condenser,  application  of,  F.  A.  Burg  316 
auxiliary,  types  of  pump  used  as  305 

Baragwanath  single  flow 294 

barometric  jet,   adjustment  and 

care 315 

Buckley  siphon,  diagram 292 

care  of 312-316 

classification 289 

cleaning 315 

cleaning,    what   determines   cost  324 
cooling    water    character,    quan- 
tity   and   source    factors   in 

selection  of 318 

counter-current 289 

double-pipe       ammonia,        with 
"Burhorn"  metallic  cooling 

tower 330 

dry-tube 300 

ejector  jet,  adjustment  and  care  315 

operation 298 

elementary  jet,  illustration 289 

volumes   of   air,    water   and 

steam,  diagram   283 

evaporation  cooling 290 

feed-water     quality     factor     in 

selection 318 

for   steam-driven   prime    mover, 

selection  of 325 

increase    in    thermal    efficiency 

due  to,  formula 281 

jet 289 

and  surface  relative  cost 323 

connected  in  parallel 296 

effect  of  bad  water 319 

ejector 289 

first    cost    less    than    that    of 

surface 323 

how    to   restore    vacuum    and 

condensing  operation 314 

low-level 289 

or  surface,  quantity  of  cooling 

water  required,  formula. . .  .   299 
power  requirements  compared 

to  surface  condenser 322 

pumping  head 320 

pumping   head    of   circulating 

pump 321 

pumps,  cost  of  maintaining . . .  323 
ratio  of  water  to  steam  fixed 

for  given  vacuum 321 

requisite  size,  formula 298 

table  of  operating  costs 326 

temperature     of     water     dis- 
charged from 300 

two  on  same  exhaust  line 296 

velocity  of  condensate  issuing 

from 299 

velocity     of     exhaust     steam 

entering 299 

with  reciprocating  engine  ....  296 
joints  and  stuffing  boxes,  index 

to  condition.  .  . .   316 

Koerting  multi-jet  ejector,  illus- 
tration      293 

leakage  of  air,  prevention  of  ....    312 

low-level  jet,  operation 295 

most  profitable  vacuum  in 285 

multi-flow 293 

illustration 294 

of  compound  condensing  engine, 

cooling  tower  data  table  ...  361 

parallel  current 289 

power  saving  due  to,  formula.  .  .  281 
power  to  remove  air  and  water 

from 284 

pump,  electric  or  steam,  type  of 
drive  for 322 


PAGE 
Condenser  pumps  .................   305 

selection  and  economics.  .  .  .   316-318 

selection  for  given  installation  .  .  316 
siphon  jet  ....................    289 

or  barometric  jet,  operation  .  .  297 
without  pump,   apparatus  for 

starting  ..................    298 

jet,  started  and  operated  with- 

out pump  ................    298 

standard  jet,  starting  ..........    295 

stopping  .................    297 

steam  ...................    277-327 

condensing  water  for  ........    329 

definition,  purpose  ..........    277 

high-temperature     cooling 

tower  test  data  table  ......   359 

saved  by  ...................    279 

saving  due  to,  graph  ........    282 

temperature  in  .............   330 

surface  ......................   289 

air-cooling  .................    290 

and  jet,  comparison  of  pump- 

ing heads  ................    320 

built-in-tube-cleaning      equip- 

ment ........  ............    324 

circulating      pump      pumping 

head  ....................    321 

discharge  from  used  for  boiler 

feed-water  ...............    388 

double  flow  ................    292 

effect  of  bad  water  ..........    319 

effectiveness      preserved      by 

exhaust-steam   separation  388 
elementary,  illustration  ......    291 

first  cost  greater  than  that  of 

jet  ......................    323 

forced-draft      cooling      tower 

with  .....................    354 

fouling  of  tubes,  result  .......    315 

heat  transference  in  .........    300 

leaky  tube  ends,  result  .......   315 

LeBlanc  ...................    311 

operation  ..................    300 

per  cent,  of  boiler  feed  returned 

to  boiler  .................    319 

power  requirements  compared 

to  jet  condenser  ..........    322 

pumping  head  ..............    320 

pumps,  cost  of  maintaining.  .  .   323 
purity    of   water   delivered   to 

feed  heater  ...............    318 

ratio  of  water  to  steam  varied 

to  suit  conditions  .....  '.  .  .  .   321 

replacement  of  tubes  ........    325 

single  flow  .................    292 

table  of  operating  costs  ......    326 

temperature  "drop"  in,  defini- 

tion .....................    304 

typical,  illustration  ..........   291 

water-cooling  ...............   290 

Westinghouse-LeBlanc  .......    311 

temperature,       vacuum       corre- 

sponding to  ..............    315 

tubes,  metals  used  for  .........   302 

tube  gland,  Worthington  stand- 


ard 

grease  removed  from 

heat  transfer  in 

sizes 
vacuum,  gages  for  measuring 

loss  while  running 

relation  to  steam  consumption  286 
Watt,  engine  with  .............   278 

Wheeler    rain  type  low-level  jet  294 
when  hot  will  not  work  ........    314 

Worthington     independent     jet, 

illustration  ............  ...    290 


302 
316 
304 
302 
284 
313 


INDEX 


429 


PAGE 

Condensing      equipment,       arrange- 
ments   ' 341 

operation,  advantages 288 

table  showing  economy 287 

work  gained  by,  diagram 280 

plant,    steam    useful    for    feed- 
water  heating 230 

plants,  vacuum  and  atmospheric 

feed- water  heaters  ........    217 

water,  centrifugal  pump  for.  .  ..    135 

for    steam  and  ammonia  con- 
densers  329 

recooling  methods 329-361 

Condensing-engine       exhaust       line, 
vacuum    trap    for    draining 

separator 404 

Conduction,  cooling  by 329 

Connecting  rods  for  deep-well  pump     88 

Contra-flow  in  economizer 265 

Convection,  cooling  by 329 

Cooling,  atmospheric,  limit  of 333 

device,    atmospheric,    efficiency, 

formula 334 

pond  area  data 342 

cooling    effect    on    condensing 

water 329 

depth 343 

diagram 337 

evaporation  rate  formula 337 

requisite  surface  area 340 

satisfactory  for  small  plants  . .    337 
simple,     requisite    area    com- 
puted     342 

spray  fountain  with 343 

tower,       atmospheric,       average 
temperature     reduction     in 

summer,  formula 356 

classification 353 

closed,  loss  of  water  less  than 

in  open 355 

closed  or  chimney-flue,  forced 

draft 354 

closed  or  chimney-flue,  natural 

draft 353 

closed      or      flue      completely 

enclosed 355 

closed,  using  either  forced  or 

natural  draft 354 

"Cooling    Tower    Company's    Cata- 
logue," formula  for  tempera- 
ture reduction  in  summer  .  .  356 
temperature      reduction 

formula : 347 

design,       distributor       and 

decks,  illustration 351 

high     temperature     natural 
draft    cooling    tower    test 

data,  table 359 

"impact"  nozzles 343 

impact  spray  nozzle 345 

installation,        spray-nozzle 

tests,  graph 346 

low      temperature      cooling 

tower  test  data,  table ....    358 
on  atmospheric  cooling  ....  333 
table  of  atmospheric  condi- 
tions      331 

computations   based    on   tests 

and  practice 355 

principles  involved 355 

construction  and  operation .  . .    350 
cooling    effect    on    condensing 

water 329 

cost 361 

fan-blower  height  required  ...   361 

for  artificial  cooling 318 

forced-draft,     for     condensers 


PAGE 

of  compound    condens- 
ing engines,  data  table.    361 
with  surface  condenser.  .  .  354 
high-temperature  natural 

draft,  for  steam  condenser, 

test  data,  table 359 

low-temperature  natural  draft, 
for     ammonia     condensers, 

test  data  table 358 

open,    loss    of    water    greater 

than  in  closed 355 

open  or  atmospheric 353 

per  cent,  of  recooling  resulting 

from  evaporation 352 

performance,  typical  data.  .  .  .   357 
proportions,    method    of   com- 
puting      360 

total  height 361 

water-loss  per  cent 357 

wood  checker  work  for 353 

Worthington,  illustration 339 

water,    character,    quantity   and 
source,  factors  in  condenser 

selection 318 

cost  of  handling 32Q 

required  for  jet  condenser.  .  .  .    299 

Corrosion  of  economizers 257 

Corrosive  liquids,  pumping 93 

Cost  of  operation  of  boiler-feed  pump  179 

Counterflow  principle 300 

Coupling,     flexible,     for     connecting 

centrifugal  pump 133 

illustration 134 

Crane   Company,    working   pressures 
for    wrought   iron  and  steel 

pipes,  data 367 

Crank-action  power  pump 75 

classification 82 

compared  with  direct-action . .      78 

data,  table 93 

for  deep  well,  classification  ...      85 

piston  speeds,  table 92 

walking-beam  type 76 

water-ends  of 80 

Crank-and-fly-wheel  pump 75 

advantages  and  disadvantages     80 

application 95 

economies 77 

hydraulic     elevator,     illustra- 
tion        78 

in  city  water  works 79 

sizes 94 

steam  pumps 79 

Cross-head  of  duplex-pump 62 

Cross-heads  secured  to  duplex-pump 

piston-rods 61 

Cup-washers    for    deep-well     pump- 
plungers  .'     98 

for     pump-plungers,     table     of 

dimensions 99 

Current-flow,  relative,  through  econo- 
mizer     265 

Cushion- valves  for  duplex-pumps  ...     64 

Cylinder  area,  definition 34 

area  of  water,  formula  for 35 

diameter  of  water  pump,  formula 

for 35 

steam,    area    to    balance    given 

water  pressure,  formula ....      37 
water,    area    to    balance    given 

steam  pressure,  formula ....      37 
Cylinder-head,      marking      striking 

point 62 


Dead-center,  avoided  by  valve-stem 

lost  motion. . .  .58 


430 


INDEX 


PAGE 

Decks    and    distributor    of    Cooling 
Tower      Company     design, 

illustration 351 

Deep-well  pump,  details  of 87-89 

pumps 84 

DeLaval   Steam    Turbine    Company, 

piston  balancing  system. ...   117 

Delivery  pipe,  friction-head  in 14 

tube  of  injector 156 

Delivery-lift  of  the  water 7 

Diagram,  indicator,   of  direct-acting 

steam-pump 25 

Diaphragm  for  pump  governor 202 

Differential      piston      for      deep-well 

pump 94 

Diffusion  vanes 109 

Direct  Separator  Company  on  sepa- 
rator economy 387 

Discharge,  actual,  of  pump 21 

level    of   condenser    higher   than 

circulating  pump 320 

line  from  injector 167 

of     piston     or     plunger     pump, 

formula 22 

pipe  of  pump,  sizes  for,  formula     18 
rate     of,      crank-action     pump, 

graph. 91 

theoretical,  of  pump 21 

Discs,  seats  of,  pump-valves 45 

Displacement      of      plunger      pump, 

formula 19 

of  pump  units  of 19 

of  reciprocating  pump  of 19 

Distribution    of    steam    from    boiler 

plant,  methods  for 371 

Distributor    and    decks    of   Cooling 
Tower      Company     design, 

illustration 351 

Double-suction  pump 114 

Draft,  artificial,  with  economizer ....   263 

available  for  economizer 273 

forced  in  cooling  tower,  advan- 
tages and  disadvantages  .  .  .   355 
in  chimney  in  inches  of  water, 

table 263 

natural,  economizer  with 262 

natural,    in    cooling    tower,  ad- 
vantages and  disadvantages  355 
Draft-pressure    drop  through  econo- 
mizer     262 

Drainage,  pumps  for 135 

Draining,       automatic,       live-steam 

separator 400 

Drive,  type  for  condenser  pumps. .  ..    322 
Driving    horse-power,    total,    steam- 
pump,   definition 29 

Driving  unit  selected  for  pump 138 

Drop,  temperature,  in  surface    con- 
densers, definition 304 

Dry-tube  condenser 300 

Dry-vacuum  pump,  illustration 79 

Duplex  boiler-feeder 204 

double-acting        power    pumps, 

application 96 

fire-pump,  illustration 27 

single-acting  pumps,  application     95 

Duplex-pump,  air-chamber  for 50 

compared  with  simplex-pumps..      64 

compound,  illustration 65 

compression-space 63 

cross-heads    secured    to    piston- 
rods 61 

for  high-pressure  service 65 

outside      adjustment      of      lost 

motion 63 

pistons,    steam-cushioned,    illus- 
tration ...  63 


PAGE 

Duplex-pump,    slow-running,    valve- 
stem  lost-motion 63 

Duty  of  pump,  basis  of  steam  con- 
sumption, formula 33 

of  pump  on  coal  basis  definition     32 
of  steam  pump,   coal  consump- 
tion, formula 32 

of  steam  pump  on  basis  of  heat 

consumed,  formula 33 

Dynamic  head  or  pressure,  definition       5 


Eclipse      exhaust-steam      separator, 

illustration 400 

jrconomizer,  advantages 272 

by-pass,  installation,  illustra- 
tion   261 

cast-iron,  wrought  iron  and 
steel,  advantages  and  dis- 
advantages    257 

coal  consumption  with,  graphs.  .  271 
compared    to    additional    boiler 

heating-surface 267 

contra-flow 265 

characteristics,  graph 266 

installed,  example 269 

corrosion  of 257 

cost  of  installation 275 

definition 210,  251 

disadvantages 272 

draft-pressure  drop  through  ....  262 
forced  draft  and  induced  draft 

fans 262 

fuel 251-275 

fuel-saving  due  to,  formula .....  270 

heat  transfer  in 269 

heat  utilized 'in 252 

heating-surface,  least  tempera- 
ture-difference for 269 

ratio  to  boiler-horsepower.  .  .  .  269 

high-  and  low-pressure 254 

independent 252 

construction 254 

illustration 253 

initial  cost 273 

inspection 274 

installation,  conditions  deter- 
mining    273 

integral 252 

integral,  high-  and  low-pressure .  255 

leakage  of  air  into 260 

minimum  temperature  difference  266 

parallel-flow 265 

characteristics,  graph 266 

ratio  of  loss  of  combustion-pas 
temperature  to  gain  of  feed- 
water  temperature 266 

relative  current-flow  through  .  . .  265 

scale  forming  in 259 

setting,   air  infiltration   through  260 

steam-generating  efficiency 271 

surfaces,  pitting  on 274 

table   of  temperatures   obtained  264 

tubes,  arrangement 255 

clean,  saving  due  to,  graph  . .  .  259 

in  staggered  rows,  illustration  256 

in  straight  rows,  illustration .  .  256 

method  of  removing  soot 257 

scale  and  sediment  in 259 

sweating 257 

tube-surfaces,  cleanliness 257 

types  of 252 

with  artificial  draft 263 

with  natural  draft 262 

Efficiencies,    relative,     boiler-feeding 

devices .  : 173 


INDEX 


431 


PAGE 

Efficiencies,  total,  steam  pumps,  table     31 
Efficiency,  hydraulic,  data  necessary 

to  determine 28 

hydraulic,  of  pump,  formula ....      27 
indicated,    reciprocating    pump, 

formula 26 

maximum      mechanical     direct- 
acting  steam-pump 30 

mechanical,  reciprocating  pump, 

definition,  formula 30 

of  pump,  causes  impairing 66 

thermal,    increase    due    to    con- 
denser, formula 281 

total,  steam-driven  pump 31 

of  pump,  definition,  formula .  .     30 
values  for  different  pumps.  ...     31 
volumetric,  of  pump,  definition, 

formula 22 

volumetric,    relation    to    pump- 
slip  22 

Ejector  jet  condenser 289 

adjustment  and  care 315 

or  jet  pump  as  condenser  aux- 
iliary    308 

priming 148 

for  pump,  illustration 146 

Elbow  in  pump  piping 6 

"Electric  Journal,"  F    A  Burg,  Ap- 
plication of  Condensers.  ...   316 
motor   drive,  centrifugal   pumps   132 
pumps,     advantages     and     dis- 
advantages, table 189 

Electrically-driven      pump,      advan- 
tages        94 

Electric-drive  for  boiler-feed  pump . .    182 
Elevator,    hydraulic,    crank-and-fly- 

wheel  pump  for 78 

pumps  for 135 

"  Elliott  Companys'    Bulletin  G"  on 
Steam-Traps,    condensation 

rates  table 412 

Engine,     condensing,     power    devel- 
oped     279 

steam  consumption  table 287 

cylinder,  water  in,  danger 386 

economy  increased  by  condenser 

creating  partial  vacuum  ....  277 
non-condensing,     power     devel- 
oped     279 

piping,     steam     separators     in, 

illustration 385 

primitive     steam,     condensation 

in 278 

reciprocating,  condenser  vacuum 

for 285 

jet  condenser  with 296 

steam  or  gasoline,   driving  cen- 
trifugal pump 133 

Watt's     condensation,     illustra- 
tion     279 

double-acting  condensing, 

illustration 279 

with  Watt  condenser. 278 

Engines,  several,  selecting  separators 

for 400 

Entrance-head,  definition 6 

Equalizer  pipes,  function 372 

Evaporation    cooling    in    condensers  290 

heat  carried  away 356 

extracted  in 340 

of  water,  cooling  effect  on  con- 
densing water 329 

rate,  cooling  pond,  formula 337 

weight,  formula 339 

Excelsior  heater  packing 246 

Exhaust     heater,     effect     on     boiler 

feeding  efficiency 174 


PAGE 

Exhaust  heater,  sound  muffled  by  ex- 
haust head 399 

steam  available,  effect   on  need 

for  economizer 273 

energy      in,       non-condensing 

plant 207 

entering  jet  condenser,  veloc- 
ity     299 

for  feed- water  heating 211 

from  engine,  heat  in 230 

heat  available  in 209 

intermittent  delivery  to  feed- 
water  heater 242 

main  piping,  average  pressure- 
drop  376 

monetary    saving    from     pre- 
heating feed-water  with. . .  .   213 

piping 363 

portion  utilized  for  feed-water 

heating 229 

separation    in    vacuum,    how 

facilitated 398 

separator,  definition 388 

separators 385-401 

supplied  by  auxiliaries 231 

temperature,      feed-water 

heater 224 

Exhaust-head,  definition 399 

pollution     of     atmosphere     pre- 
vented by 399 

purpose 399 

Expansion  in   pipe  of  given  length, 
pipe     required     for     bend, 

formula 378 

in  steam  separators 389 

linear,  coefficient  of 378 

in  steam  pipes,  strains  due  to  377 
in     steel     and     wrought-iron 

steam  pipes,  formula 378 

stresses  in  piping  systems,  taken 

up  by  bends 370 


Fan  pumps 117 

Fan-blower      height      required      for 

cooling  tower 361 

Fair  banks- Morse     multi-stage     cen- 
trifugal   pump,    illustration  113 
Farnsworth    boiler    feeder,    illustra- 
tion  204 

Feed-pump,  boiler,  connection  to 
feed-water  heater,  illustra- 
tion    3 

boiler,       mechanically-driven, 

illustration 182 

capacity,  forcing  boilers 197 

constant-speed,  water-relief 

valve  on 201 

locate  close  to  heater 245 

motors  for  driving 184 

suction  connection 2 

Feed-water,  chemical  treatment 318 

cold,  boiler  strain  due  to 160 

waste  of  fuel  by 207 

cost  of  treating 319 

formula  for  determining  require- 
ments     195 

fuel  saving  due  to  preheating.  . .    211 

heater 207-249 

as  a  purifier 238 

as  protective  measure 228 

atmospheric 215 

back-pressure  valve 220 

Blake-Knowles 233 

centrifugal    pump   in    connec- 
tion with 136 


432 


INDEX 


PAGE 
Feed-water      heaters,       classification 

table 215 

cleaned  regularly 248 

closed 211 

advantages    and    disadvan- 
tages      244 

classification 238 

installation  and  operation .  .   247 

National  coil  type 238 

piping 247 

safety  valve 247 

tubes  in 240 

type,       table       of       general 

data 236,  237 

Cochrane,  illustration 226 

open  induction 219 

connection     to     boiler     feed- 
pump         3 

counter-current,  diagram 239 

diagram 239 

double  installation 244 

economies 211 

exhaust  steam,  definition 210 

temperature 224 

function 224 

heat  transmission  table 243 

heater  horsepower  rating 243 

heating-surface,  formula 242 

horizontal  closed 220 

induced  or  draw  heaters 220 

induction,  piping  of 223 

induction-type    open,    piping, 

illustration 222 

installing    primary    and    sec- 
ondary, illustration 216 

intermittent    delivery    of    ex- 
haust steam 242 

open 210 

advantages       and       disad- 
vantages     244 

explanation 225  * 

installation 245 

size  of  shell  required 234 

type,  table  of  general  data  235 

typical  installation 221 

operation 246 

pan  or  tray  area  required ....    232 

piping 2 

arrangements 220 

pressure 215 

primary 215 

and      secondary,      operated 

alternately 218 

application 216 

purity   of   water   delivered   to 

by  surface  condenser 318 

reason  used 207 

savers  of  coal 214 

secondary 215 

open  or  closed  type 217 

selection 238 

steam  condensed  by,  formula.    231 

steam-tube 238 

application 241 

illustration 239 

Stilwell    through-type,    piping 

arrangement 223 

through        or        thoroughfare   . 

heaters 220 

used  with  injector 160 

vacuum 215 

and    atmospheric,    in.  con- 
densing plants 217 

water-tube 238 

heating  equipment,  classes 210 

saving      due      to      computed 
graphically 213 


PAGE 

Feed-water,  impurities 227 

live-steam  heaters  and   purifiers  248 

oily 228 

per    cent,    returned    to    boiler  in 

surface  condenser 319 

preheating   with   exhaust  steam, 

monetary  saving  resulting ..   213 
quality,      factor     in     condenser 

selection 318 

quality  for  economizer 273 

requirements,  power  plant,  basis 

of  steam  consumption 196 

of      power      plant,       estima- 
tion     194-197 

temperature  gain,  in  economizer, 
ratio  to  loss  of  combustion- 
gas  temperature 266 

temperature  for  economizer  ....   273 

raised,  formula 228 

with  surface-condenser  used  re- 
peatedly    288 

Filtering  material,  packing 246 

Filtration  in  open  heater 225 

Fire-pump,  duplex,  illustration 27 

pumps  for 135 

underwriters' 59 

Fire-insurance    underwriters,    pumps 

required  by 65 

Fisher  pump  governor . 200 

illustration 198 

Fiske,  R.  A.,  on  turbine  drives 133 

on    centrifugal   pump    advan- 
tages      134 

single  impeller  pump 120 

squirrel-cage     induction     motor, 

centrifugal  pump  drives". .  ..    132 
Fittings,      extra      heavy      cast-steel, 

illuslratipn 365 

malleable    iron,    illustration  364 
frictional-resistance  in,  table.  ...      15 
to  water-flow  through  .....      14 
low-pressure    cast-iron,    illustra- 
tion      365 

right-angled,  in  steam  pipes, 
pressure-drop  due  to,  for- 
mula   377 

standard    cast-iron,    illustration  364 
malleable      iron,      illustration  364 
Flanges,      companion,      methods      of 

securing  to  pipe  ends 369 

cost,  according  to  material 369 

methods    of    attaching    to    pipe 

ends 368 

Flexibility  of  boiler  operation  due  to 

economizer 272 

Float  connections  in  open  heater. .  . .    246 
Flow  through  valves  in  piping,  fric- 
tion due  to 6 

Flue-gas     temperature     for     econo- 
mizer    273 

Foot-valve  in  suction  line 142 

leakage,  effect  of 67 

priming    ejector    use    with    and 

without 148 

Force,  centrifugal,  definition 101 

illustration 102 

Forcing  boilers,  feed-pump  capacity  197 

Formulas,  reciprocating  pump 34 

Foster,     D.    E.,    steam    pipe    sizes, 

graph 374 

Foundation     bolts     for     centrifugal 

pump 140 

for  a  centrifugal  pump 140 

Fountain,     see     also     Spray     foun- 
tains     347-350 

spray,  with  cooling  pond 343 

Friction  due  to  flow 6 


INDEX 


433 


PAGE 

Friction  in  pipe,  effect  on  pump-suc- 
tion         1 

in  straight  pipes 5 

mechanical,  in  pump  mechanism     30 

of  liquid  entering 6 

of  water  in  straight  pipes 9 

Friction-head,  definition 4 

in  centrifugal  pump 130 

of  water  in  cast-iron  pipe,  table 

of 12 

in  straight  pipes,  table  of.  ...      10 

on  a  pump 5 

on     a     pump,     definition,     for- 
mula          6 

Frictional  resistances,  what  included     25 

Fuel  economizers 251-275 

see  also  Economizers. 
heat  of,  in  gases  of  combustion, 

table 251 

saved     by     heating    feed-water, 

graph 213 

savine  due  to  economizer 272 

formula 270 

due  to  preheating  feed- water.    211 
waste  of  by  cold  feed-water ....    207 

Fulton  governor,  illustration 199 

on     direct-acting     boiler-feed 

pump 178 

pump  governor,  explanation.  .  .  .    198 

Funnel,  non-splash 167 

Funneled  inlet-orifice,  example 17 


Gage  glass,  shielded  from  steam- 
temperature  fluctuations, 

illustration 401 

water,  on  steam  separator 401 

Gage-pressure,  net,  how  determined .      28 

Gallons  per  minute,  formula  for 34 

Gas  in  feed-water 228 

Gases,  exit-temperature  in  econo- 
mizer   269 

relative  flow  in  economizer 265 

Gear  drive  preferable  to  belt  drive  185 
Gearing,  inspection  in  economizer.  .  .    274 
"Gebhardt's     Steam     Power    Plant 
Engineering,"       advantages 
and   disadvantages   of   open 
and        closed        feed-water 

heaters 244 

table    of    engine    econ- 
omy     287 

list  of  heater  manufacturers.  ...    215 

on  heat  transfer 243 

"Gillette  and   Dana,"   on  separator 

costs 401 

Gland,  condenser  tube 302 

"  Goulds  Manufacturing  Company's 
Catalogue,"  on  boiler-feed- 
pump sizes 193 

on    motor    ratings,    deep-well 

pump 97 

centrifugal  pump  material 101 

open-well  pump,   illustration  ...      83 

pump  capacities 193 

rotary  pump 152 

table 2 

table    of    boiler-feed    centrifugal 

pump  capacities 194 

thrust  bearing 121 

Governor,  boiler-feed  pump,  on 
direct-acting  steam 

pumps 198-201 

28 


PAGE 
Governor  for  turbine-driven  pumps, 

discharge-pressure,  control  .    133 
pump,  horizontal  type  .........    199 

on   turbine-driven    centrifugal 
pumps  ...................    20  1 

troubles  of  .................    201 

Gravity    apparatus    or    return    traps 

for  boiler-feed  ............    171 

Grease  on  outside  of  condenser  tubes, 

how  removed  .............    316 

Greene  Economizer  Company  boiler 

heating-surface  charts  .....    268 

table  of  economizer  temperatures 

obtained  .................    264 

showing  heat  of  fuel  in  gases 
of  combustion  ............    251 

temperature  charts  ............    269 

Green  economizer,  illustration  ......   253 

Gridiron  separator  ................    389 

operating  principle  ..........    394 

Ground-area  required,  spray-fountain 

ponds  ...................    350 

Guide     bearing,     centrifugal     pump, 

illustration  ...............    122 

coupling  for  deep-well  pumps.  .  .      88 
vanes  ........................    109 


H 


Hampton     Mills,     East     Hampton, 
Massachusetts,    installation 
of  economizer  .............    253 

Hancock  inspirator,  illustration  .....    157 

Hanger,  counter-balancing,  for  steam 

piping  ...................    380 

Hangers,  plain,  for  steam  piping  ....   380 

"  Harding   and    Willard,  Mechanical 
Equipment    of    Buildings," 
data  on  feed-water  heaters  235 
"  Harrison  Safety  Boiler  Works  Cata- 
log,"    condenser      economy 
graph  ....................    286 

condenser  saving  ............   282 

separators  ....................    387 

Head,  dynamic,   or  pressure,  defini- 

tion .....................        5 

exhaust-,  definition  ............   399 

friction,  definition  .............        4 

on  a  pump  .  .  ...............        5 

inlet     static,     for     boiler     feed- 

pumps ...................        4 

maximum,  impellers  for  ........    119 

measured,  definition  ...........        5 

in  pump  operation  ..........        7 

necessary  to  overcome  frictional 

resistance  .  .  .............  .      14 

of     water,     converting     to     unit 

pressure,  formula  .........        4 

pressure,  and  speed  of  impeller, 

centrifugal  pump  ..........    108 

static,  of  a  liquid,  illustration.  .  .        4 
of  fluid  column,  definition.  ....    3 

total  measured,  on  pump,  defini- 

tion, formula  .............        9 

on      pump,    definition,    f  o  r  - 
mula  ....................        9 

pumped  against  .............    106 

velocity,  definition  ............        4 

Headers,  duplicate  main,  boiler  con- 

nection to  ................   372 

Heat     absorbed     by     water    in     in- 

jector ....................    162 

abstracted  from  steam  by  cooling 
water  in  surface-condenser, 
formula  .......  .  ..........    302 

balance,       automatic       exhaust 

steam  ...................    181 


434 


INDEX 


PAGE 

Heat  extracted  in  evaporation 340 

insulating  for  closed  heater 246 

latent,       of      vaporization,       in 

water  cooling 329 

loss,  excessive,  from  steam  pipes, 

how  prevented 382 

from   bare   steam    pipes,    con- 
densation due  to,  formula  381 
and  insulated  steam  pipe. . .   381 

in  non-condensing  plant 209 

of  fuel  in  gases  of  combustion, 

table 251 

pump-duty    on    basis  of,  for- 
mula        33 

saved  by  preheating  boiler  feed- 
water 212 

transfer     coefficient     in     closed 

heaters 242 

in  economizer 269 

transference     coefficient,     value 
effected     by     conditions    in 

condenser  tubes 304 

in  surface  condensers 300 

condenser,    table    of    coeffi- 
cients     304 

transmission,  feed-water  heaters, 

table 243 

Heater,  closed,  non-condensing  plant, 

duplex  boiler  feeder  in 204 

exhaust,  effect  on  boiler  feeding 

efficiency 174 

feed-water,         see         Feed-water 

heater 3 

closed,    advantages    and    dis- 
advantages     244 

open,    advantages  and  disad- 
vantages      244 

steam  trap  for,  illustration .  .  .   405 

used  with  injector 160 

formula  for  raising  temperature 

of  feed-water 228 

horsepower     rating,     feed-water 

heaters 213 

live-steam,  for  feed-water 248 

Heaters,  feed-water 207-249 

Heater-tube,  corrugated 241 

Heating  coils,  steam  traps  for 409 

exhaust-steam,  heat  balance  for  182 
system,       gravity,       condensate 

from 227 

Heating-surface,     additional     boiler, 

compared  to  economizer .  .  .    267 
boiler,     temperature     difference 

for 268 

economizer,    least    temperature- 
difference  for 269 

feed-water  heater,  formula 242 

Hershey  Chocolate  Company's  Plant, 

illustration xii 

Holly   steam-loop   for   draining   high 

pressure  piping,  illustration  382 
Hoppes    feed-water    heater,    illustra- 
tion    221 

live-steam  heater 232 

purifier    installed,    illustration  248 
reverse-current     exhaust     steam 

separator,  illustration 390 

Horsepower    delivered    by    injector  163 
heater,  feed-water  heater  rating  243 

of  pump 25 

required  for  pumping,  formula.  .      32 
total      driving,      steam      pump, 

definition 29 

water,      developed      by      pump, 

formula 25 

water    or    hydraulic,    indicated, 

definition  formula 29 


PAGE 
Hot  water    handled  with  centrifugal 

pumps 137 

pumping 2 

returns,       feeding       of      boilers 

with 202 

Humidity,  relative  average,  table.  .  .    331 
corresponding     to     wet-     and 
dry-bulb     temperature    dif- 
ferences, table 335 

definition 332 

effect  on  water  cooling 330 

of  atmospheric  air,  how  deter- 
mined      332 

Hydraulic  efficiency,  data  necessary 

to  determine 28 

of  pump,  formula 27 

or   water   horsepower   developed 

by  pump,  formula 25 

packing  for  pumps 42 


Impact  or  baffle-plate  separator 389 

Impeller,     centrifugal     pump,     theo- 
retical speed 106 

closed-type,  illustration 119 

enclosed 119 

for  maximum  heads 119 

forces  which  tend  to  unbalance  114 

Jaeger  method  of  balancing.  ...  115 

open 117 

r.p.m.  and  velocity  of  periphery, 

formula 108 

speed,  effect  of  change  in 129 

of,    and    pressure    head,    cen- 
trifugal pump 108 

Impurities  scale  inside  of  boiler 207 

scale-forming 248 

Indicator  card,  steam  pump 28 

diagram,     direct-acting     steam- 
pump 25 

Infiltration  of  air  through  economizer 

setting 260 

Injectors 155-170 

advantages 159 

application 173 

in  absence  of  heater 173 

approximate  equation 161 

as  a  pump,  inefficient 172 

becoming  hot 172 

breaks 170 

capacities     and     weights,     table  166 

different  types,  applications.  .  .  .  160 

disadvantages 160 

double-tube 157 

efficiency 174 

elementary,  illustration 156 

essential  parts 156 

factors  influencing  performance  .  163 

failure  to  lift  water 169 

fed  from  overhead  tank,  illustra- 
tion   168 

for  boiler-feed '.  .  .  .  171 

horsepower  delivered  by 163 

inspirator       type,      piped       for 

boiler  feeding,  illustration.  .  173 

installating 166 

lifting 157 

limitations  for  boiler-feeding. . . .  172 

measure  of  economy 162 

Metropolitan  Model-0 159 

non-lifting 157 

not  delivering  water  to  boiler. .  .  169 

operating  and  starting 169 

Penberthy    automatic,    illustra- 

.tion 158 

piping  of,  illustration 167 


INDEX 


435 


PAGE 

Injectors,  positive,  operation 159 

principle  of,  illustration 155 

selection 165 

self  adjusting 158 

single-tube    automatic,    restart- 
ing   157 

special,  for  exhaust  steam 155 

steam  at  overflow 170 

steam  pressure  for 172 

suction-pipe  strainer 168 

theory 156 

troubles 169 

type  for  given  service 165 

Inlet    static    heads    for    boiler    feed- 
pumps   4 

Inlet-orifice,  funneled,  example 17 

Inspection  of  economizers 274 

Inspirator,  Hancock,  illustration.  .  .  .  157 
type   injector,    piped    for    boiler 

feeding,  illustration 173 

Instruments,  economizer  fitted  with  .  259 

Insulation  for  steam  pipes 381 

Intake-pipe  to  suction-well 9 

International   Text   Book   Company, 

table  of  engine  economy .  .  .  287 

Iron,  cast,  for  steam  piping 363 

malleable,  for  steam  piping 363 

wrought,  for  steam  piping 363 

Irrigation,  pumps  for 135 


Jacobus,    D.    C.,    9n    economies    of 

boiler  feeding 174 

Jaeger  method  of  balancing  impeller  115 
of  impeller  balancing,  illustra- 
tion     116 

Jet    condenser,    see    also    Condenser, 

jet , 289 

elementary,    volumes    of    air, 

water   and   steam,    diagram  283 
high-vacuum,     with     turbine, 

illustration 305 

Johns  electrically-driven  pump 188 

Joints,    condenser,    index    to    condi- 
tion     316 

corrugated    expansion,    illustra- 
tion      378 

double-slip    expansion,    illustra- 
tion     378 

double   swing   or  swivel,   taking 
up  expansion  in  pipe  lines, 

illustration 377 

expansion  slip- 377 

flanged,  in  steam  piping 368 

screwed,  in  steam  piping 368 

steam-piping,  types  used  in 368 

Josse,  Professor,  tests  on  heat  trans- 
ference coefficient 304 


K 


Kansas  City  Light  and  Power  Com- 
pany economizer  installa- 
tion   254 

Kelley,  H.  H.,  on  condensers 312 

Kent's  "Mechanical  Engineers' 
Pocketbook  "  on  economies 

of  boiler  feeding 174 

on  open  heater  tray  area 232 

Kieley  expansion  steam-trap,  illustra- 
tion    404 

pump  governor,  illustration  ....    200 

Kinghorn  pump  valve,  illustra- 
tion   46 

Kneass'  "  Practice  and  Theory  of  the 

Injector  " 164 


PAGE 
Koerting  multi-jet  ejector  condenser, 

illustration 293 

multi-spray  nozzle 344 

roof    space    for    spray    cooling, 

illustration 346 

Kroeshell         Brothers         Company, 

Chicago,  power  plant 208 


Lap- weld,  strength  of 367 

Lap-welded  pipes 365 

Law  of  freely  falling  bodies 105 

Leakage  due  to  cold  water  in  boiler  207 
of  air  in  condenser,   prevention 

of 312 

Leathers,     pump-plunger,     mold    for 

forming. 99 

LeBlanc  surface  condenser 311 

Lift,    suction,    of    centrifugal    pump, 

how  measured 8 

Liner  for  piston  pump 42 

Liquid,    corrosive,    centrifugal   pump 

for  moving 151 

pumping 93 

entering,  friction 6 

volatile  non-corrosive,  pumping.      93 
Live-steam  heater  for  feed- water.  .  .  .    248 

piping 363 

purifier  for  feed-water 248 

separation,  economy. 386 

purposes  of 386 

separators 385-401 

definition 385 

efficiencies  table 396 

Load,    character   of,    for   economizer  273 
on  pump,  practice  in  determin- 
ing       28 

Loew  absorption  exhaust-steam  sepa- 
rator, illustration 395 

Loop  header  system  for  steam  piping  371 
Holly,  returning  condensation  in 

steam  piping  to  boilers 383 

Loss  by  incorrect  valve-stem  adjust- 
ment       61 

due  to  inefficient  boiler-feeding 

pumps 178 

heat,      excessive,      from     steam 

pipes,  how  prevented 382 

from  bare  and  insulated  steam 

pipes,  table 381 

steam    pipes,    condensation 

due  to,  formula 381 

in  non-condensing  plant 209 

hydraulic,  how  obtained 28 

in  pump  tests 27 

of  pump,  definition 26 

in  power  plant,  chart 252 

of  water,  greater  in  open  than  in 

closed  cooling  tower 355 

Lostrmotion,  valve-stem,  function  of 

in  steam-pumps 58 

Louvres,    Burhorn   sheet-metal  cool- 
ing tower 354 

in  cooling  tower 352 

Low  pressure  steam,  table  of  prop- 
erties    301 

Lubrication  of  engine  cylinder  when 

wet  steam  is  used 386 


M 


Magnesia  heat-insulating  for  closed 

heater 246 

Main  pipe  with  branches,  size,  for- 
mula   376 


436 


INDEX 


PAGE 

Make-up  water,  cooling  effect 342 

definition 356 

Management,  pump 66 

Marck  expansion  steam-trap  for 
feed-water  heater,  illustra- 
tion    405 

Marks'  "Mechanical  Engineer's 
Handbook  "  on  boiler  feed- 
ing equipment 173 

on  centrifugal  pump  character- 
istics     126 

on  steam  pipe  insulation 381 

Masher  centrifugal  horizontal 
steam  separator,  illustra- 
tion   391 

Massachusetts      pump,      illustration   101 
Mechanical     drive     for     boiler-feed 

pump 182 

efficiency     reciprocating     pump, 

definition,  formula 30 

Measured  head,  definition 5 

"Mechanical   and    Electrical   Cost 

Data,"  on  separator  costs..   401 
"  Mechanical  Engineers'  Handbook," 

on  steam  pipe  insulation.  .  .    381 

"  Mechanical  Refrigeration  " 332 

Mercury,    inches    of,    conversion    to 

pounds  per  square  inch ....      29 

vacuum  gage 284 

Mesh  separator 389 

operating  principle 393 

Metropolitan  Model-O  injector 159 

Midvale    Machine    Company,    boiler 

feed  test 188 

Mine  drainage,  pumps  for 135 

Mitchell-Tappen    System,    high-tem- 
perature natural  draft  cool- 
ing tower  test  data  table.  .  .    359 
low   temperature    cooling    tower 

test  data,  table 358 

Mixing  chamber,  Worthington  cool- 
ing tower 352 

M  off  at  feed- water  heater 224 

Moisture  carried  from  boiler  as  spray 

or  bulk  of  water 386 

in  steam  from  boiler 385 

Momentum  in  steam  separators 389 

Motor,     adjustable-speed,     for    feed 

pumps 185 

aligning  with  vertical  pump  shaft  122 
constant-speed,     on     boiler-feed 

pump 185 

direct-current 132 

for  driving  feed  pumps 184 

rating      for      deep-well      pump, 

formula 97 

slip-ring     induction,      for     cen- 
trifugal pumps 132 

squirrel-cage      induction,      cen- 
trifugal pump  drives,  R.  A. 

Fiske 132 

varying-speed,  for  feed  pumps .  .    185 
Motor-driven       centrifugal       pumps 

speed  variation 151 

Motor-generator   in    automatic   heat 

balance 182 

Mover,  steam-driven  prime,  selection 

of  condenser  for 325 

Muntz    metal    for    condenser    tubes  302 


N 


Nason      bucket-float      steam      trap, 

illustration 403 

Nation  coil  heater 241 

Newcomen's      condensation      engine  278 


PAGE 

Non-condensing  and  condensing 
operation,  steam  consump- 
tion with 280 

plant,    duplex    boiler    feeder    in 

closed  heater 204 

energy  in  exhaust  steam 207 

exhaust  steam  available 229 

location  of  separator 399 

steam     useful     in     feed-water 

heating 230 

Non-return  trap,  definition 403 

Nozzles,  see  also  Spray  noz- 
zles   344-348 

for     spray     fountain     size     and 

number 348 

impact,     Cooling     Tower    Com- 
pany      343 

in  spray  fountain,  spacing 348 

intermingled    spray    from,    illus- 
tration      345 

pump     discharge,     pressure     at, 

illustration 26 

single  spray,  illustration 344 

spray,    impact,    Cooling    Tower 

Company 345 

Schutte  and  Koerting  cooling 

pond 338 

steam,  of  injector 156 


Oil  in  boiler  feed-water 211 

removed  from  feed- water 227 

Oil-drip    connection    in    open   heater  246 

Oil-eliminators,  definition 388 

Oil-separators,  cost 401 

definition 388 

part  of  heater 224 

Oiling  machinery 69 

Operating     costs,     jet     and     surface 

condensers,  table 326 

Operation,  condensing,  advantages .  .    288 
Overflows  of  injector 157 


Packing,  cutting  down 43 

effect  on  pump  suction 1 

of  governor  valve  stem 201 

pump  rods  and  stems 69 

water-piston 42 

Pan  area  required,   open  feed-water 

heater 232 

Pans  of  open   heater,   removed  and 

cleaned 247 

Parallel-flow  in  economizer 265 

Parsons  vacuum  augmenter,  illus- 
tration   307 

"  Peele's  Mining  Engineers  Hand- 
book," table  of  pump  effi- 
ciencies    31 

Penberthy  automatic  injector,  illus- 
tration   158 

Pipe  anchorage,  illustration 380 

butt-welded,    method    of    form- 
ing    365 

capacities  for  saturated  or  super- 
heated steam,  graph 374 

cast-iron,      table      of      friction- 
heads 12 

connections   for   injectors,   table  166 

delivery,  friction-head  in 14 

obstruction 169 

discharge,  of  steam-trap,  check- 
valves  in 413 

sizes  for,  formula 18 

velocity  through,  formula.  ...      24 
double  extra  heavy  grade 363 


INDEX 


437 


PAGE 

Pipe,  extra  heavy  cast  iron,  for  steam- 
piping  systems 364 

cast-steel,   for  steam-piping 

systems 364 

grade 363 

malleable   iron,    for   steam- 
piping  systems 364 

fittings,  grades  used  in  steam- 
piping  systems 364 

for  induction  heater 221 

for  mixing  chamber 352 

for    steam-power    plant,    grades 

of,  table 368 

hanger,  illustration _.    380 

in  steam-piping  systems  classi- 
fied according  to  construc- 
tion    365 

lap-welded,  method  of  form- 
ing   366 

preferable  to  butt-welded  for 

steam  piping 367 

length  necessary  for  bend  to  take 
up  expansion  in  pipe  of 

given  length,  formula 378 

lengths  equivalent  in  resistance 

to  fittings,  table  of 15 

lines,  double  swing  joint  taking 

up  expansion  in 377 

uncovered,  steam  condensa- 
tion rate  in,  table 412 

low-pressure  cast-iron,  for  steam- 
piping  systems 364 

main,  size  formula. 376 

pump      suction,      with      square 

orifice *••>        6 

size     for    reciprocating     engine, 

formula 375 

necessary  to  deliver  steam  at 

given  rate,  formula 375 

saturated        or       superheated 

steam,  graph 374 

spiral-riveted  steel 366 

standard  cast-iron  for  steam- 
piping  systems 364 

grade 363 

malleable  iron  for  steam- 
piping  systems 364 

steam,  bare  and  insulated,  heat 

losses  from,  table 381 

condensation  due  to  heat  loss, 

formula  and  table 381 

excessive  heat  loss,  how  pre- 
vented   382 

insulation , 381 

linear     expansion     producing 

strains  in 377 

pressure  drop  due  to  globe 
valves  and  right-angled  fit- 
tings, formula 377 

sizes  determined  graphically. .    373 

thickness  of  covering 382 

steel  and  wrought,  sizes 364 

and  wrought-iron,  steam, 
linear  expansion  in,  for- 
mula    378 

grades 363 

trade  meaning 367 

straight,  friction  of  water  in.  ...        9 

riveted  steel 366 

wrought-iron    or    steel,    table 

of  friction  heads  in 10 

suction,    for   condenser,   strainer 

in 313 

sizes  for.  formula 18 

table  of  friction-heads  in 10 

vacuum,  centrifugal  pump  suc- 
tion    .145 


PAGE 

Pipe   vent,   connecting  high-pressure 
trap          with         apparatus 

drained 413 

induction  heater 246 

vibration,     devices    to    prevent 

transmission,  illustration..  .    379 

welded  wrought-iron 367 

steel 367 

wrought-iron,  inside  and  outside 

diameters 363 

trade  meaning 367 

Pipe-bend  facilitating  steam  flow  in 

systems 369 

minimum  length  of  tangent.  ...   371 

radius 371 

pressure-drop  produced  by 377 

radii  of .    370 

resistance     to     steam     flow     in 

piping  system  decreased  by  369 
standard,  for  piping  systems. . .  .    370 

Pipe-ends,  flanges  attached  to 368 

Pipe-fittings,  friction 6 

Piping,   discharge,   centrifugal  pump  145 
engine,     steam     separators     in, 

illustration 385 

exhaust-steam 363 

main,  average  pressure-drop.  .    376 

for  boiler  feeding 171 

friction    due    to    flow    through 

valves 6 

high-pressure,   Holly  steam-loop 

for      draining,      illustration  382 

live-steam 363 

main  steam,  unit  system,  illus- 
tration     373 

of  steam  trap 412 

pump,  turn  made  with  elbow ...        6 
with      long-radius      bend       6 
sharp  turn,  made  with  plugged 

tee 6 

steam,  floor  stands  for 380 

high  pressure,  condensation 
returned  to  boilers  with 

Holly  loop 383 

loop      header      or      duplicate 

headers  for 371 

materials  for 363 

of  power  plants 363-383 

of  power  plant,  two  separate 

systems 363 

plain  hangers  for 380 

screwed     and     flanged    joints  368 

single  header  system 371 

supporting  devices  for 380 

types  of  joints  used 368 

unit  group  for 371 

vibration     due     to     pulsating 

steam-flow 379 

wall-brackets  for 380 

suction,  centrifugal  pump 141 

system,  expansion  stresses  taken 

up  by  bends 370 

quantity  of  condensation- 
water  to  be  trapped,  for- 
mula   411 

steam,  grades  of  pipe  fittings 

used 364 

Piston   balancing   system,    De  Laval 

Steam     Turbine     Company  117 

pump,  effective  area 20 

requisite    diameter    for    water 

end,  formula 22 

volume  swept  by,  illustra- 
tion   19 

speed,        crank-action        pump, 

table 92 


438 


INDEX 


PAGE 

Piston,    steam,    direct-acting    steam- 
pump,  formula  for  diameter.     23 
effect     of     vacuum,     illustra- 
tion     277 

Piston-pumps,  discharge-heads 42 

illustration 41 

Piston-speed,  high,  relation  to  pump- 
slip 21 

Fitting    on    surfaces    of    economizer  274 
Plant,   non-condensing,   heating   sys- 
tem,    boiler-feeding    equip- 
ment for 182 

Plunger,   pump,   dimensions   of   cup- 
washers,  table 99 

effective  area 20 

inside-packed,  illustration.  ...      20 
requisite    diameter    for    water 

end,  formula 22 

rods,     deep-well     pump,     head- 
pressure    equivalents,    table     98 
Plunger-pump,      belt-driven      single 

acting -. 78 

discharge  heads 42 

outside  end-packed 41 

pump,  illustration 40 

Plunger-valve,  deep-well  pump 88 

Pond,     cooling,      condensing     water 

cooled  by 329 

simple,  requisite  area  com- 
puted   342 

spray-fountain,       ground       area 

required 350 

Pond-area,  requisite  total 341 

Pot-valves,  illustration 45 

Pot- valve  type  of  pump  valve 47 

Power  developed  by  condensing  and 

non-condensing  engine 279 

horsepower  of  pump 25 

required  for  pumping,  for- 
mula   32 

total    driving,    steam    pump, 

definition 29 

water  or  hydraulic,  indicated, 

definition  formula 29 

"Power,"  on  glass  water-gages 401 

on  pipe  trade  meanings 367 

power  house  drawing  from.  .  .      iv 
Power    plant    auxiliaries    and    acces- 
sories, illustration 208 

condensing  steam,  non-con- 
densing steam-driven  aux- 
iliaries...  > .  .  180 

"  Power  Plant  Engineering,"  on  boiler 

feeding 205 

on  economizer  cost 275 

R.  A.  Fiske,  on  centrifugal  pump 

advantages 134 

Power   plant,    estimating   feed-water 

requirements 194-197 

feed-water  requirements  on 
basis  of  steam  consump- 
tion   196 

losses  in,  chart 252 

steam,    grades     of    pipe    for, 

table 308 

piping 363-383 

Power    pump,    advantages    and    dis- 
advantages, table 189 

belt-driven 82 

chain-driven 82 

driven  by  gearing 82 

duplex 82 

rates  of  discharge 90-92 

simplex 82 

triplex 82 

^quired  to  drive  centrifugal 
pump  at  any  speed 130 


PAGE 

Power  pump,  to  operate  spray  foun- 
tain     350 

requirements  of  jet  compared  to 

surface  condenser 322 

saving    due    to    condenser,    for- 
mula     281 

to  remove  air  and   water  from 

condenser 284 

Power-factor-correcting  synchro- 
nous-motor-driven centrifu- 
gal pumps 132 

"Practical  Engineer,"  on  cost  of  cool- 
ing tower 361 

"Practical  Heat" 329 

Preheater  in  Badenhausen  boiler.  .  .  .    255 
Pressure,     absolute,     in     condenser, 

formula 285 

barometric,   effect   on   condenser  285 
draft,   in   chimney,   in  inches  of 

water,  table 263 

drop,  average,  in  exhaust-steam 

main  piping 376 

due  to  globe  valves  and  right- 
angled     fittings     in     steam 

pipes,  formula 377 

in    steam    mains,    allowed    in 

practice 376 

prevented         by         receiver- 
separator  388 

produced  by  gate  valves  and 

pipe  bends 377 

saturated       or       superheated 

steam  graph 374 

friction,  definition 4 

head,  definition 5 

head  and  speed  of  impeller,  cen- 
trifugal pump,  formula.  .  .  .    108 

due  to  in  the  vessel 8 

inlet,  for  boiler  feed-pumps 4 

measured,  definition 5 

of  exhaust,  selection  of  separator 

affected  by 400 

producing  velocity  in  a  pipe.  ...        5 
pump  intake,   at  different  tem- 
peratures, graph 2 

unit,   converting  head  of  water 

to,  formula 4 

velocity,  definition 4 

water-vapor,  in  air,  how  deter- 
mined     336 

working,  increased  by  condenser  277 
wrought  iron  and  steel  pipe, 

data 367 

Priming  centrifugal  pumps 146 

methods  of 147 

ejector 148 

use    with    and    without    foot- 
valve 148 

excessive,  prevented  by  receiver 

separator 388 

Priming-pump  for  centrifugal  pump.    147 

with  low  suction-lift 149 

Proper  pump,  selection  of 93-97 

Proportions  of  cooling  tower,  method 

of  computing.  .  . 360 

Psychrometer,    sling,     to    determine 

relative  humidity. 332 

Pumps,  see  also  Duplex-pumps  and 

simplex-pumps. 

see     also     Plunger-pumps     and 
piston-pumps. 

actual  discharge  of 21 

work  of,  definition 24 

air   below   suction-valves,   effect 

of 66 

between  suction  and  discharge 
decks,  effect  of 67 


INDEX 


439 


PAGE 
Pump,    Alberger   hurling-water    air, 

illustration 307 

and  receiver,  combined 202 

and  suction  pipe,  passage  of 
water  through,  as  a  hydrau- 
lic loss 27 

apparatus  for  replenishing  air- 
chamber,  illustration 49 

artesian- well 85 

belt-driven  single-acting,  illus- 
tration    77 

blowing  out  steam  cylinders.  ...  68 

boiler-feed,  cost  of  operation 179 

economical 179 

electric-drive  for 182 

mechanical  drive  for 182 

mechanically-driven,   capacity 

of 186 

constant-speed  economy.  185 

motor-driven,  illustration 177 

motor-    or    power-driven,     in 

non-condensing  plant 179 

steam-driven  reciprocating,  in 

every  plant 179 

steam-piston 65 

table  of  capacities 193 

Burnham     compound     simplex, 

illustration 64 

calculations 1-37 

capacity  at  low  speed 66 

causes  impairing  efficiency 66 

centrifugal,  see  also  Centrifugal 
pumps. 

advantages  of 134 

and  rotary 101-153 

as  condenser  auxiliary 307 

boiler-feed,  illustration 177 

characteristics,  graph 125 

efficiency  for  boiler  feeding.  .  .  192 

methods  of  priming 147 

suction  lift  of,  how  measured .  8 

circulatory,  for  condenser 305 

lower  than  discharge  level  of 

condenser 320 

classes  of,  used  with  condenser  305 

cleaning  out  steam  piping 68 

combination     high-service     and 

low-service  belt-driven 77 

compared    to    steam    traps    for 

boiler  feeding 205 

compound  condensing,  duty  and 

steam  consumption,  table .  .  72 

deep-well 87 

direct-acting 65 

starting 67 

condensate  for  condenser 305 

condenser,  electric  or  steam,  type 

of  drive  for 322 

crank-action,  see  also  Crank- 
action  pumps 75—99 

power 75 

suction  and  discharge,  graphs  91 

crank-and-fly-wheel 75 

application 95 

operation 76 

deep-well,  Chippewa,  illustra- 
tion    86 

details  of 87-89 

illustration 84 

motor  rating  for,  formula ....  97 
plunger     rods,     head-pressure 

equivalents,  table 98 

direct-acting     boiler-feed,     with 

Fulton  governor 178 

feed,  efficiency 174 

simple,  duty  and  steam  con- 
sumption, table 71 


PAGE 
Pumps,  direct-acting,  steam,  see  also 

Steam  pumps,  direct-acting3Q-73 

as  condenser  auxiliary 305 

boiler-feed      pump      gover- 
nors     198-201 

for  boiler-feed  service,  selec- 
tion of 65 

indicator  diagram 25 

maximum    mechanical    effi- 
ciency    30 

requisite  steam-piston  diam- 
eter, formula 23 

steam-driven,    application    for 

boiler  feeding 180 

discharge  pipe,  sizes  for,  for- 
mula    18 

discharging  into  reservoir 17 

displacement  units  of 19 

double-acting,  computations. ...  34 

duplex  crank-and-fly-wheel. .  .  76 

simplex,  application 95 

suction,  illustration 1 

triplex,  illustration 81 

double-suction,  balancing  of.  ...  117 

dry-vacuum,  illustration.  ......  79 

or  air,  for  condenser 305 

Wheeler 310 

duplex,  adjustment  of  steam- 
valve 57 

compared     with     simplex- 

pumps 64 

cross-heads 62 

definition 53 

displacement  of,  example 20 

double-acting  power,  applica- 
tion    96 

fire,  illustration 27 

illustration 54 

requisite     length     for    steam- 
valve  rod 57 

steam- valve 55 

electric,  power  and  steam,  ad- 
vantages and  disadvantages, 

table 189 

electrically-driven,  advantages. .  94 
failure   to    catch    water    due   to 
leakage  of  valves  in  suction 

chamber 67 

feed,  saving  by  substituting 
electrically-driven  for  steam- 
driven 187 

feed-water,  table  of  applications  190 

for  boiler-feed 171 

for  liquids  other  than  water.  ...  93 

for  water-service  in  buildings ...  59 

friction  in  valves 6 

frictional  resistance  offered  by 
internal  passages  and 

valves 14 

friction-head,  definition,  for- 
mula    6 

friction-head  on 5 

Goulds,  open-well,  illustra- 
tion    83 

governor,  horizontal,  illustration  199 
on   turbine-driven    centrifugal 

pumps 201 

troubles  of 201 

governor-controlled  duplex,  illus- 
tration   59 

high        pressure,        air-chamber 

charging-apparatus  for 51 

horizontal  double-acting  suction, 

arrangement  of  valves 47 

hot-well 305 

hurling-water,  as  condenser  aux- 
iliary   307 


440 


INDEX 


PAGE 

Pumps,  hydraulic  efficiency    of,  for- 
mula    27 

losses,  definition 26 

or  water  horsepower  developed 

by,  formula 25 

hydro-centrifugal,    as    condenser 

auxiliary 307 

incorrect    adjustment    of    valve- 
stem  as  source  of  loss 61 

inefficient     boiler-feeding     losses 

due  to 178 

inlets  connected  to  two  or  three 

sources 171 

inside-packed,  illustration 41 

installing 68-69 

jet,  as  condenser  auxiliary 308 

load     on,     practice     in     deter- 
mining   28 

management 66 

manufacturer,     data     furnished 

to 137 

Massachusetts,  illustration .....  101 
mechanically-driven,    for    boiler 

feeding 176 

modern  applications 97 

motor-driven,  for  boiler  feeding.  176 

net  work  of,  definition,  formula.  24 

new,  running .  66 

of  jet  condenser,  cost  of  main- 
taining    323 

of    surface    condenser,    cost    of 

maintaining 323 

oiling 69 

operation,  measured  heads  in.  . .  7 
outside    plunger-pump,    illustra- 
tion    40 

piston,      high      vacuum,      how 

secured 310 

illustration 41 

or      plunger,      discharge      of, 

formula 22 

volume     swept     by,     illustra- 
tion    19 

plunger,    displacement    of,    for- 
mula   .-  •  •  •  ; 19 

inside-packed,  illustration.  ...  20 

or  piston,  effective  area 20 

power,         see         also         Power 

pumps 82-97 

power        feed,         boiler-feeding 

efficiency 176 

power-driven,  total  efficiency.  .  .  31 

power-plant,  suction  piping 40 

pressure     at     discharge     nozzle, 

illustration.. 26 

priming,  explanation 67 

raising  water  when  empty 67 

reciprocating     compound,     duty 
and      steam      consumption, 

table 72 

discharge  velocity,  formula. .  .  24 

displacement  of 19 

formulas 34 

indicated  efficiency,  formula. .  26 

indicator  diagram 25 

mechanical    efficiency,    defini- 
tion, formula 30 

net  suction-lift,  definition ....  2 

rods  and  stems,  packing  of 69 

rotary,  action  of 152 

advantages      and      disadvan- 
tages   153 

application 153 

definition,  illustration 152 

rotative  or  crank-action,  as  con- 
denser auxiliary 305 

run  continuously 69 


PAGE 

Pumps,  runners,  life  of 323 

set   below  suction  supply,   illus- 
tration   139 

simplex,   compared   with  duplex 

pumps 64 

definition 53 

illustration 54 

length     of     stroke,      explana- 
tion    57 

single  impeller,  R.  A.  Fiske 120 

single-acting     duplex,      applica- 
tion    95 

triplex,  diagram 92 

illustration 81 

snifter      for      replenishing      air- 
chamber,  illustration 50 

stand-by,  direct-acting 192 

starting 70 

steam,  duty  of,  definition 32 

end  warmed  up 70 

total  efficiencies,  table 31 

steam-driven  crank-and-fly- 

wheel,  illustration 75 

for  boiler  feeding 176 

total  efficiency 31 

stopping 70 

submerged-piston 47 

suction  lifts  at  various  altitudes, 

table 2 

suction-pipe,  funneled  end,  illus- 
tration    6 

sizes  for,  formula 18 

tests,  hydraulic  losses 27 

theoretical  discharge  of 21 

water  lift  at  different  tempera- 
tures, graph 2 

total  driving  horsepower,  defini- 
tion    29 

efficiency,       definition,       for- 
mula    30 

values 31 

head,  definition,  formula 9 

measured  head,  definition,  for- 
mula    9 

triplex     double-acting,     applica- 
tion    96 

single-acting    power,    applica- 
tion    96 

turbine,  definition 109 

types    used    as    condenser    aux- 
iliaries   305 

vacuum-chamber,  function  of. . .  51 
Vaile-Kimes  single-acting  deep- 
well ......  94 

vertical  duplex,  boiler  feed,  illus- 
tration  -.-.•••  55 

volumetric  efficiency,  definition, 

formula 22 

volute,  definition 109 

water       run        through        when 

stopped 67 

water-level  in  air-chamber 50 

wet  vacuum,  for  condenser 305 

Wheeler-Edwards  combined 

condensate  and  air 309 

work,  in  horsepower 26 

Pump-duty,  basis  of  heat  consumed, 

formula 33 

of    steam    consumption,    for- 
mula   33 

Pump-piston,     canvas-packed     illus- 
tration   43 

canvas  packing  rings  for 43 

metal  packed,  illustration 42 

water-packed,  illustration 42 

Pump-plungers,  deep-well 86 

dimensions  of  cup-washers,  table  99 


INDEX 


441 


PAGE 

Pump-plungers  for  deep-well,  illustra- 
tion    87 

leather  cup  for  packing 88 

or  piston,  requisite  diameter  for 

water  end,  formula 22 

Pump-slip   affected    by    high   piston- 
speed 21 

average  values 21 

definition 21 

explanation 21 

negative 21 

percentage,  formula 21 

relation  to  volumetric  efficiency  22 
Pump-suction,  effectiveness  in  lifting 

water 1 

Pump-valve,  ball,  illustration 45 

bronze  disk  type,  illustration ...  44 

conical-seated,  illustration 44 

effective  area  of  opening 48 

flat-seated,  illustration 44 

in  power-plant  pumps 43 

Kinghorn,  illustration 46 

seats  of  metal-disc 45 

securing  into  valve  deck 46 

stems  and  piston  rods,  causes  of 

scoring 70 

used   for   clear   liquids,    illustra- 
tion    46 

Pump-work,  useful,  illustration 7 

Pumping  engines 79 

head,    jet-condenser    circulating 

pump 321 

surface-condenser     circulating 

pump 321 

horsepower  required,  formula. .  .  32 

of  hot  water 2 

unit,  selection  of 138 

Purification    of   feed- water   by    open 

heater 224 

Purifier,  feed-water  heater  as 238 

live-steam,  for  feed-water 248 

R 

Radiation,  cooling  by 329 

Radojet,      Wheeler     two-stage      air 

pump,  illustration 308 

Rain  type,  Wheeler  low-level  jet  con- 
denser  .  .  .  294 

Rayne's  formula  for  computing  pipe 

length  required  for  bend.  .  .  378 

Receiver  and  pump,  combined 202 

Receiver-separator,     Cochrane    hori- 
zontal, illustration 387 

definition 387 

Reciprocating  engine  condenser  prac- 
tice    285 

pipe  size  for,  formula 375 

pumps,  compared  to  centrifugal 

pumps 134 

compound    condensing,    duty 
and     steam      consumption, 

table 72 

duty   and   steam   consump- 
tion, table 72 

displacement  of 19 

formulas 34 

indicated  efficiency,  formula. .  26 
mechanical    efficiency,    defini- 
tion, formula 30 

net  suction-lift,  definition.  ...  2 

principle  of,  illustration 40 

simple,  duty  and  steam  con- 
sumption, table 71 

Recooling,    atmospheric,    of    conden- 
sing water 329 

condensing  water,  methods.  329-361 


PAGE 

Recooling,  effected  in  cooling  tower, 
per  cent  resulting  from  evap- 
oration    352 

in    spray    fountains    conditions 

affecting 344 

system,  devices  for  bringing  air 

and  water  into  contact 337 

Relative  humidity,  definition 332 

effect  on  recooling 345 

Resistance  due  to  water  friction 9 

to  steam  flow  in  piping  system 

decreased  by  pipe  bends .  .  .   369 

Return  trap 202 

boiler-feeding 203 

definition 403 

or  gravity  apparatus  for  boiler- 
feed 171 

Reverse-current  separator 389 

operating  principle 389 

Rings,  canvas  packing  for  pump  pis- 
ton    43 

metallic,  packing  for  pumps.  ...      42 
snap,  water-piston  packed  with, 

illustration 39 

pipes 365 

steel  pipe,  spiral 366 

straight .  .  .   366 

Rods  and  stems   of  pump,   packing 

of 69 

steam-valve,       duplex-pumps, 

illustration 58 

piston,  causes  of  scoring .      70 

Roof,    power-house,    spray-fountains 

on 350 

Roof-space  for  spray  cooling,  illustra- 
tion   346 

Rotary  pump,  definition,  illustra- 
tion   152 

Royal  Technical  School,  Charlotten- 
burg,  tests  on  heat  transfer- 
ence coefficient 304 

S 

Safety  valve,  closed  feed-water  heater  247 

inspection  in  economizer 274 

Saving  by  substituting  electrically- 
driven  for  steam-driven  feed 

pump 187 

due  to  feed-water  heating  com- 
puted graphically 213 

monetary,  resulting  from  pre- 
heating feed-water  with  ex- 
haust steam 213 

power  due  to  condenser,  for- 
mula   281 

Scale  forming  in  economizers . 259 

inside  boiler  from  impurities. . .  .    207 

Scale-forming  impurities 248 

Schutte     and     Koerting     Company, 

spray-nozzle  capacities  table  348 
table  of  spray-fountain  data. .  .  .   349 
double     spraying    system,     dia- 
gram   338 

straight-tube  closed  heater .       .  .    240 

cooling  guarantees 347' 

Scraper,  economizer-tube 257 

tube,  power  expended 258 

Screens  for  mixing  chamber 352 

Sealing  surfaces,  fit  between 119 

Seats  of  metal-disc  pump-valves ...      45 

Sediment  in  economizer-tubes 259 

Sellers'  injector,  self-adjusting  num- 
ber 8,  performance,  graph  164 

Restarting  Injector 165 

Separating-plate,  Bundy  steam  sepa- 
rator, illustration 394 


442 


INDEX 


PAOE 
Separation  efficiency  and  velocity  of 

steam  flow  graph 397 

exhaust-steam,  economy 388 

in  vacuum,  how  facilitated .  .  .    398 

purposes 388 

live-steam,  economy 386 

purposes  of 385 

maximum  efficiency  attainable. .   395 

Separator,  absorption 389 

and  appurtenances  for  efficiency 

test,  illustration 397 

Austin       reverse-current       live- 
steam,  illustration 390 

centrifugal 389 

operating  principle 390 

efficiency  affected  by  velocity  of 

steam-current 396 

exhaust-steam,  definition 388 

Loew  absorption 395 

proper  location  for 399 

selection  of 400 

gridiron 389 

Hoppes  reverse-current  exhaust- 
steam,  illustration 390 

Harrison    Safety    Boiler    Works  387 

impact  or  baffle-plate 389 

live-  and  exhaust-steam. .  .  .    385-401 
in  engine  piping,  illustration. .   385 
live-steam,     Austin     baffle-plate 

angle,  illustration 391 

definition 385 

drained  automatically 400 

efficiencies  table 396 

efficiency  of,  formula 397 

on  basis   of  steam  quality, 

formula 398 

high-pressure  steam  trap  for, 

illustration 405 

operation  and  structure 387 

proper  location  for 399 

selection  of 400 

storage  capacity 387 

Stratt9n    centrifugal  horizon- 
tal, illustration 391 

with  large  well 387 

location,  selection  affected  by. . .   400 

mesh 389 

oil,  cost 401 

for  feed-water 227 

of  feed-water  heater,  illustra- 
tion      225 

receiver-,  defini^n 387 

excessive  primin  prevented  by  388 
pressure  drop  prevented  by ...   388 

steam-storage  capacity 388 

vibration  prevented  by 388 

reverse-current 389 

operating  principle 389 

steam,  Bundy  gridiron,  illustra- 
tion      394 

classification 389 

cost 401 

functions    of    corrugated    sur- 
faces     393 

Masher     horizontal     centrifu- 
gal,   illustration 391 

physical  phenomena  involved  389 

size 400 

Swartwout    centrifugal,    illus- 
tration     391 

Sweet        vertical,         illustra- 
tion     393 

with  glass  water-gages 401 

vacuum  trap  for  draining 404 

Welderon      reverse-current      re- 
ceiver-, illustration 390 

well,  function 387 


PAGE 
Setting,    economizer,    leakage   of   air 

into 260 

inspection  in  economizer 274 

Sewage,       centrifugal       pumps       for 

moving 151 

fumping  plants,  pumps  for. ....    135 
,    size    required    for    open    feed- 
water  heater 234 

Side-suction  pump 114 

Side-thrust  in  centrifugal  pump 114 

Simplex  double-acting  pumps,  appli- 
cation       95 

pump,  air-chamber  for 50 

as  vacuum-  and  air-pumps  ...  65 
compared  with  duplex-pumps  64 
single-acting,  rates  of  suction 

and  discharge,  graph 90 

steam-valve 55 

Single     header     system     for     steam 

piping 371 

Single-suction  pump 114 

Siphon  jet  condenser 289 

Sling     psychrometer     to     determine 

relative  humidity 332 

Slip,  pump,  definition,  explanation.  .      21 
relation      to       high       piston- 
speed 21 

Smallwood,  Julian,  "Mechanical 

Laboratory  Methods  " 161 

Snifter  for  air-chamber  of  pump,  illus- 
tration       50 

Speed    variation,    motor-driven    cen- 
trifugal pumps 151 

Spray  cooling  and  condenser-outfits, 

performance  guarantees. . .  .   347 
constant,    proper    value,    pre- 
determination     347 

effect  on  condensing  water.  .  .    329 
roof  space  for,  illustration.  .  .  .    346 
fountain,  conditions  affecting  re- 
cooling  344 

in     connection     with     cooling 

ponds 343 

installations,   table  of  related 

data 349 

nozzles,  size  and  number 348 

spacing 348 

on  power  house  roof 350 

ponds,  ground-area  required. .  350 
power  required  to  operate.  .  .  .  350 

nozzle,  Badger 344 

capacities,  table 348 

impact,  Cooling  Tower  Com- 
pany     345 

installation,    temperature    re- 
duction effected  by,  formula  347 

tests,  graph 346 

pond  for  artificial  cooling 318 

with     Cooling     Tower     Com- 
pany's    impact     nozzles...   343 
Springs,  pump  discharge  valves,  pres- 
sure to   overcome   reaction, 

as  a  hydraulic  loss 27 

Stand-by  pump,  direct-acting 192 

Standoipe,  column  of  water  in  a ....        3 

illustration 4 

Stands,  floor,  for  steam  piping 380 

Static  head  of  fluid  column,  definition    .  3 
Steam     and     oil    separators,    Direct 
Separator      Company,      on 

separator  economy .   387 

at  injector  overflow 170 

automatic  exhaust,  heat  balance  181 
boilers,    apparatus    for    feeding 

water 171 

condensation  rate  in  uncovered 

pipe  lines,  table 412 


INDEX 


443 


PAGE 
Steam,  condensed  by  open  feed-water 

heater,  formula 231 

condenser,     see    also     Condenser 

steam. 277-327 

condensing  water  for 329 

consumption,  duty  of  pump  on 

basis  of,  formula 33 

power  plant  feed-water  re- 
quirements based  on 196 

relation  to  condenser  vacuum  286 
with  condensing  and  non-con- 
densing operation 280 

crank-and-fly- wheel  pumps 79 

end,  pump,  warmed  up 70 

energy  conserved  by  separation.    386 
exhaust,  see  also  Exhaust  steam. 
energy      in,       non  condensing 

plant 207 

heat  in 209 

main  piping,  average  pressure- 
drop  376 

separating  for  heating  system  388 

separators 385-401 

from  boiler,  moisture  in 385 

heat  abstracted  from  by  cooling 
water  in  surface  condenser, 

formula 302 

heating,  low-pressure 180 

how  saved  by  condenser 279 

in  condensing  plant,   useful  for 

feed-water  heating 230 

line  tapping  for  injector 166 

live,  separators 385-401 

low    pressure,     table    of    prop- 
erties    301 

main,  floor  stand  supporting. . . .    380 
pressure-drop  allowed  in  prac- 
tice    376 

suspending  and  counterbal- 
ancing expansion  loops 380 

wall-bracket  supporting 380 

net  thermal  value  diminished  by 

moisture 386 

nozzle  of  injector 156 

pipe  size  necessary  to  deliver  at 

given  rate,  formula 375 

Steam     piping,     see      also     Piping, 

steam 363-383 

of  power  plants 363-383 

power  plant,  grades  of  pipe  for, 

table 368 

pumps,    advantages    and    disad- 
vantages, table 189 

allowable  velocity  in  water- 
piping  40 

direct-acting 39-73 

classification 41,  53 

for  boiler-feed  service,  selec- 
tion of 65 

hydraulic      pressure,      illus- 
tration       45 

outside  center-packed,  illustra- 
tion    20 

reciprocating  double-acting. . .      39 
vacuum-chamber       connected 

to,  illustration 52 

valve-stem  lost-motion,  func- 
tion of 58 

with  outside  end-packed 
plungers,  water  end,  illustra- 
tion    20 

quality,   efficiency  of  live-steam 

separator  based  on 398 

saturated    or  superheated,  pres- 
sure-drop,    pipe    sizes    and 

capacities  for,  graph 374 

saving  due  to  condenser,  graph  .    282 


PAGE 
Steam  separators,  see  also  Separators, 

steam 385-401 

cost 401 

supplied  by  boiler  plant,  methods 

of  distributing 371 

useful     in     feed-water     heating, 

non-condensing  plant 230 

volume  required  for  discharge  of 

return  trap 406 

water  in,  turbines  damaged  by  386 
wet,    effect    on    engine    cylinder 

lubrication 386 

loss  of  turbine  efficiency  due 

to 387 

reasons  for  separating 386 

Steam-bound,  pumps  becoming 4 

Steam-current  velocity  through  sep- 
arator, efficiency  affected 

by 396 

Steam-flow  in  piping  system,  resist- 
ance decreased  by  pipe- 
bends 369 

pulsating,    causing    vibration   in 

steam  piping 379 

velocity  and  separation  efficiency 

graph 397 

velocities  in  practice 375 

Steam-gage  pressure  to  balanpce 
water-gage  pressure,  for- 
mula    36 

Steam-loop,  Holly,  for  draining  high- 
pressure  piping,  illustration  382 
Steam-piston      diameter,      requisite, 
direct-acting    steam    pump, 

formula ..      23 

Steam-storage       capacity,       receiver 

separator 388 

Steam-temperature  fluctuations, 

shielding   glass    water-gages 

from 401 

Steam      traps,      see      also      Traps, 

steam 403-413 

"Bulletin"  G,  Elliott  Company, 

condensation  rates,  table 412 

"Catechism,  Swendeman's,"  ca- 
pacities       and        dimensions, 

table 411 

compared    to   pumps   for   boiler 

feeding 205 

separator  drained  by 400 

Steam-valve,  adjustment  of,  duplex- 
pump  57 

of  duplex-pump 55 

rod,       duplex-pump,       requisite 

length 57 

Steel,  cast,  for  steam  piping 303 

mild,  for  steam  piping 363 

pipe  and  wrought  iron,  working 

pressures,  data 367 

trade  meaning 367 

Stems  and   rods   of  pumps,   packing 

of 69 

steam-valve,       duplex-pumps, 

illustration 58 

pump-valve,    causes    of    scoring     70 

Stilwell  feed-water  heater 217 

through-type  feed-water  heater, 

piping  arrangemen 223 

Soot,     methods    of    removing    from 

economizer-tubes 257 

Soot-blower,  economizer 258 

Soot-blowing  system,  steam-con- 
sumption of 258 

Soot-pit,  insp'ection  in  economizer.  .  .    274 

Soot-scraper,  illustration 258 

inspection  in  economizer 274 

Storage  of  feed-water  in  open  heater  225 


444 


INDEX 


PAGE 

Strainer,  foot-valve  with,  illustra- 
tion    142 

in     circulating     water    suction 

pipe 313 

in   trap-inlet   connections,   func- 
tion     412 

injector  suction-pipe 168 

pump       suction-pipe,       illustra- 
tion          6 

Strains,  avoided  by  preheating 173 

Stratton  centrifugal  horizontal  live- 
steam  separator,  illustra- 
tion   391 

Stresses,  boiler,  diminished  by  use  of 

economizer 272 

Striking-points  equally  spaced,  illus- 
tration   G2 

Stroke,    length    in    inches,    formula 

for 35 

Strokes  per  minute,  formula  for 36 

Stuffing-box,     condenser,     index     to 

condition 163 

Stuffing-boxes   of  centrifugal  pumps, 

packing 151 

Sturtevant,  B.  F.  Company,  econ- 
omizer draft  fans 262 

functions,  diagram 251 

economizer,     header    and     tube 

construction 256 

Suction-lift,     net,     of     reciprocating 

pump,  definition 2 

of  centrifugal  pump 139 

how  measured 8 

of   pump    at    various    altitudes, 

table 2 

of  the  water 7 

practical  maximum 1 

pump  with  low,  priming 149 

Suction  line,  enlarged,  connected  to 

centrifugal  pump. . . . ' 143 

imperfectly  laid,  illustration .  .        8 
independent,  centrifugal 

pumps 142 

nozzle,  pump,  pressure  in,  illus- 
tration       40 

pipe  of  pump,  sizes  for,  formula     18 
pump,     funneled     end,    illustra- 
tion          6 

rate     of,     crank-action     pump, 

graph 91 

Suction-supply,  distant 9 

Suction  well  for  pump  supply 9 

illustration 8 

Sump  for  single  centrifugal  pump .  . .    143 
Surface  condenser,  see  also  Condenser, 

surface 289 

coefficients    of    heat    transfer- 
ence, table 304 

feed-water  used  repeatedly ...   288 
heat  abstracted  from  steam  by 

cooling  water  formula 302 

tubes  and  tube-sheets 302 

Surface  condensing  plant  location  of 

separator 399 

corrugated,  in  steam  separator, 

functions 393 

external,  and  internal,  inspection 

in  economizer 274 

water-cooling,  required  in  surface 

condenser,  formula 303 

Swartwout  centrifugal  steam  sepa- 
rator, illustration 391 

Sweating  economizer-tubes 257 

Sweet  mesh,  exhaust-head,  'illustra- 
tion    399 

vertical  steam  separator,  illus- 
tration    393 


PAGE 

Swendeman's,  "A  Steam-Trap  Cate- 
chism," capacities  and  di- 
mensions table 411 

Symbols,  list xii 

Synchronous-motor-driven  centrifu- 
gal pumps,  power-factor- 
correcting 132 


Tee,  plugged,  for  sharp  turn  in  pump 

piping 6 

Temperature,  breaking,  of  injector.  .    164 
"drop"    in    surface    condensers, 

definition 304 

effect  on  pump  intake  pressures, 

graph 2 

of  atmospheric  air,   how   deter- 
mined      332 

of  water,  effect  on  pump  suction       1 
reduction     effected     by     spray- 
nozzle    installation    formula  347 
average,    effected    by    cooling 

tower  in  summer 356 

wet  and  dry-bulb,  table 331 

Temperature-difference      for      boiler 

heating-surface 268 

least,    for    economizer    heating- 
surface  269 

wet-      and      dry-bulb,      relative* 
humidities  corresponding  to, 

table 335 

Terminal  difference,  definition 320 

jet  condenser 320 

Terry     Steam     Turbine     Company, 

pump  test 136 

Testing  centrifugal  pump,  illustration 

and  formulas 123 

Tests,  pump,  hydraulic  losses 27 

Thermometer,  dry-bulb 332 

economizer  fitted  with 259 

wet-bulb 332 

Thrust-bearing      for      vertical      sub- 
merged    centrifugal     pump   152 
Goulds      Manufacturing      Com- 
pany     121 

Tile-tubing  for  mixing  chamber 352 

Tilting     or     dumping     steam     trap, 

counterweighted 406 

Total  efficiency  of  pump,  definition, 

formula 30 

Tower,     cooling,     see     also     Cooling 

tower 339-361 

Burhorn  metallic 330 

condensing    water    cooled    by  329 
Trap,       ball-float,       valve-operating 

mechanism,  illustration 407 

Bundy,  illustration 203 

continuous-discharge 405 

discharge     pipes,      check-valves 

in 413 

draining  discharge  pipes 412 

expansion,  location 410 

inlet    connections,    strainers    in, 

function 412 

intermittent-discharge 405 

return 202 

coal  saving  effected  by 408 

operating  principle 405 

volume  of  steam  required  for 

discharge 406 

steam 403-4 13 

capacity 410 

care  of 413 

classification  according  to  dis- 
charge     405 

to  operation 403 


INDEX 


445 


PAGE 
Trap,  steam,  compared  to  pumps  for 

boiler  feeding 205 

definition 403 

dimensions      and      capacities, 

table 411 

expansion,       for       feed-water 

heater 405 

external  by-pass 412 

for  heating  coils 409 

for  live  steam  separators,  illus- 
tration     405 

high-  or  low-pressure,  location 

for 409 

Kieley  expansion,  illustration.   404 
methods  of  detecting  leaks.  .  .    413 

non-return,  economy 408 

piping 412 

return  and  non-return 403 

economy 407 

proper  location  for.  .......    408 

temperature-operated,    limita- 
tions      409 

vacuum,   for  draining  separator  404 
vent-pipe        connecting         with 

apparatus  drained 413 

Trapping-sheet    in    steam    separator  393 
Tray  area  required,  open  feed-water 

heater 232 

Trays,       perforated,       for       mixing 

chamber 352 

Triplex  double-acting  pumps,  appli- 
cation    96 

single-acting       power       pumps, 

application 96 

Tubes      and      tube-sheets,      surface 

condenser 302 

condenser,  fouling  of,  result.  ...    315 
economizer,  scale  and  sediment 

in 259 

in  closed  feed-water  heater 240 

surface-condenser,  replacement  325 
Tube-cleaner,  Worthington  hydraulic  324 
Tube-cleaning  equipment,  built-in, 

surface  condenser 324 

Tube-sheets,  surface  condenser 302 

Tube-surface  required  in  surface  con- 
denser, formula 303 

Tube-surfaces,  economizer,  cleanli- 
ness   257 

Turk  piston-packing,  illustration...  .      43 

Turbine  condenser  practice 286 

vacuum  for 285 

damage  due  to  water  in  steam .  .   386 
for     centrifugal     pump,      water 

rate 193 

pumps,  definition 109 

steam,  as  pump  drive,  condens- 
ing     133 

driving  centrifugal  pump 133 

with      high-vacuum      jet      con- 
denser     305 

Turbo-generator,  Westinghouse- 
LeBlanc  surface  condenser 
with 311 


U 


Uniflow-engine  condenser  practice. . .    286 

Underwriters,  fire-insurance,  pumps 

required  by 65 

Underwriters'  fire-pump,  illustra- 
tion    59 

Unit  system  main  steam  piping, 

illustration 373 

Unit-group  arrangement  for  steam 

piping 372 


PAGE 


Vacuum    augmenter,    Parsons,   illus- 
tration     307 

condenser,  loss  while  running.  .  .    313 
relation    to    steam    consump- 
tion      286 

conversion  inches  of  mercury  to 

pounds  per  square  inch ....      29 
corresponding  to  condenser  tem- 
perature      315 

exhaust-steam  separation  in,  how 

facilitated 398 

gage,  illustration 284 

high,    with    piston    pumps,    how 

secured 310 

in  jet  condenser,  how  restored.  .  314 
most  profitable,  in  condenser .  .  .  285 
partial,  in  condenser,  increased 

engine  economy 277 

Vacuum-chamber        connected        to 

steam-pump,         illustration     52 
height  of  water  in,  illustration .  .      52 

in  pumps,  function  of 51 

special,  illustration 52 

Vacuum-pump,  simplex  pump  as. ...      65 
Vaile-Kimes    single-acting    deep-well 

pump 94 

Valve  arranged  above  pump-barrel, 

illustration 47 

back-pressure,  feed-water  heater  220 
decks,  securing  pump  valve  seats 

in 46 

direct-acting  simplex  steam- 
pump,  illustration 56 

discharge,      closed,      centrifugal 

pump  run  with :  .  .    149 

effect     of    tightness     on     pump 

suction 1 

exhaust-relief,  back-pressure,  in 

water  piping 246 

flat  disc,  area  of  opening 49 

flat-faced  bronze  poppet 47 

wing  poppet,  illustration 46 

flat-seated  pump,  illustration  ...  44 
foot,  with  strainer,  illustration. .  142 
frictional  resistance  in  pump  due 

to 14 

table 15 

to  water-flow  through 14 

gate  or  check,  in  pump  discharge 

line 67 

pressure-drop  produced  by  ...    377 
globe,  in  steam  pipes,  pressure- 
drop  due  to,  formula 377 

in   piping,   friction   due   to   flow 

through 6 

inside-operated,  simplex  steam- 
pump,  illustration 56 

leakage,  failure  to  catch  water.  .  67 
leaky,  in  steam  boiler-feed 

pumps 192 

of  horizontal  double-acting 
suction-pumps,  arrange- 
ment    47 

of  power-plant  pumps 43 

pump,  see  also  Pump-valves. . .   43—47 
•discharge,    pressure    to    over- 
come reaction  of  springs,  as 

a  hydraulic  loss 27 

friction  in 6 

rubber,  for  low  pressure 46 

seat  of  governor 202 

uneven  wear  in 68 

side  by  side  above  pump-barrel, 

illustration 48 

steam,  of  simplex-pumps 55 


446 


INDEX 


PAGE 
Valve,  steam-pump,  wear  causing  lost 

motion 68 

steam-thrown 55 

suction,     air    below,     effect     on 

pump 66 

uneven  wear  in 68 

water-relief,    on    constant-speed 

feed  pump 201 

Valve-discs,       rubber       composition, 

hardness 45 

Valve-operating  mechanism  of  Ameri- 
can   ball-float    steam    trap, 

illustration 407 

Valve-orifice  of  steam  trap,  area.  .  .  .    408 

Valve-stem  lost-motion,  correct 62 

in  steam-pumps,  function  of. .      58 
slow-running        duplex-pumps     63 

of  governor,  packing 201 

pump,    incorrect   adjustment   as 

source  of  loss 61 

rigid,      connection     in      duplex- 
pump 63 

Vanes,  diffusion 109 

guide 109 

Vapor   pressure    of   saturated    water 

vapor,  graph 336 

water,  in  air,  weight,  how  deter- 
mined      336 

pressure    in    air,    how    deter- 
mined      336 

Velocity,  allowable  in  water-piping  of 

direct-acting  steam  pump .  .      40 

steam-flow,  in  practice 375 

head,  definition 4 

why  neglected 5 

of  freely  falling  body,  formulas  105 
of     periphery     and     r.p.m.     of 

impeller,  formula 108 

current  through  separator,    effi- 
ciency effected  by 396 

flow  and  separation  efficiency 

graph 397 

in  separator  selection 400 

of  water  in  cast-iron  pipe,  table 

of ... 12 

in  pipes,  table  of 10 

pressure  producing  in  a  pipe.  ...        5 
through  reciprocating  pump  dis- 
charge pipe,  formula 24 

Vent-pipe    connecting    high-pressure 
trap          with          apparatus 

drained 413 

induction  heater 246 

Vertical  pump,  aligning  motor  with  122 

bearings  in 121 

Vessel,  head  due  to  pressure  in 8 

Vibration    in    steam    piping    due    to 

pulsating  steam-flow 379 

pipe,   devices  to  prevent  trans- 
mission, illustration 379 

prevented  by  receiver-separator  388 
Voltage,    steady,     electrically-driven 

centrifugal  pumps 151 

Volume  of  water  ana  friction  head  in 

centrifugal  pump 130 

swept  by  pump  piston,  illustra- 
tion       19 

Volumetric  efficiency  of  pump,  defini- 
tion, formula 22 

relation  to  pump-slip 22 

Volute  pumps,  definition 109 

Vulcan  soot  blower 258 

W 

Wainwright  closed  feed-water  heater, 

copper  corrugated  tube ....    240 


PAGE 

Washers,     cup,   for  deep-well  pump- 
plungers  98 

Waste-valve  of  injector 157 

Water  and  air,  power  to  remove  from 

condenser 284 

apparatus     for     feeding     steam 

boilers 171 

area     of    in     cylinder,     formula 

for.  .  .  . 35 

bad,  effect  in  jet  condenser 319 

on  surface  condenser 319 

column    of,    converting    to    unit 

pressure,  formula 4 

condensing,  see  also  Condensing 
water. 

recooling  methods 329-361 

cooling,  character,  quantity  and 
source  factors  in  condenser 

selection 318 

cost  of  handling 320 

devices  for  bringing  into  contact 

with  air  in  recooling  system  337 
discharged   from   jet    condenser, 

temperature 300 

heads,  high,  effect  on  condenser 

water  requirement 322 

height  raised  by  pump  suction.  .        1 

hot,  pumping 2 

"make-up,"  cooling  effect 342 

definition 356 

passage  through  suction  pipe 
and  pump,  as  a  hydraulic 

loss 27 

pounds    pumped    per    pound    of 

steam 162 

pumped  per  pound  of  steam 
decreases  with  steam  pres- 
sure. .- 164 

quantity  of,  delivered  by  cen- 
trifugal pump 107 

raised  to  height 112 

ratio  to  steam,  jet  and  surface 

condensers 321 

relative  flow  in  economizer 265 

slugs,  dangers 386 

vapor  in  air,  weight,  how  deter- 
mined    336 

pressure    in    air,    how    deter- 
mined      336 

saturated,  pressure  graph  ....    336 
weight  of,  in  one  cubic  foot  of 

air,  graph 336 

Water-cooling  in  surface  condenser.  .    290 
surface      required      in      surface 

condenser,  formula 303 

Water-end,  crank-action  pumps 80 

of    direct-acting    steam    pump, 

illustration 39 

Water-gage,  glass,  on  steam  separator  401 
pressure  to  balance  steam-gage 

pressure,  formula 36 

Water-horse-power       developed     by 

pump,  formula 25 

indicated,  definition  formula. .      29 
Water-loss,  cooling  tower,  per  cent .  .    357 
greater  in   open  than  in  closed 

cooling  tower 355 

Water-piston,  packed  with  snap  rings, 

illustration 39 

packing 42 

Water-relief  valve  on  constant-speed 

feed  pump 201 

Wearing  rings 115 

illustration 116 

Welderon     reverse-current     receiver- 
separator,  illustration 390 

Well,  deep-well  pump  for 84 


INDEX 


447 


PAGE 

Well,  driven,  centrifugal  pump 144 

or   sump    for    single    centrifugal 

pump 143 

suction,  illustration. . . . 
Westinghouse-LeBianc    surface    con- 
denser .  . 


turbine  with,  illustration 


..   311 
..   306 

Wheeler  cooling  tower 351 

dry-vacuum  pump 310 

rain    type     low-level     jet     con- 
denser    294 

two-stage  "Radojet"  air-pump, 

illustration 308 

Wheeler-Balcke          natural         draft 

cooling  tower 353 

Wheeler-Edwards  condensate  and  air 

pump 309 

Whitlock  closed  heater,  manifold 241 

Wind  velocities,  table 331 

Wind-break  around  spray  fountain.  .    350 

Wing- valves  in  high-pressure  pumps     45 

Work,  actual,  of  pump,  definition ...      24 

gained  by  condensing  operation, 

diagram 280 


PAGE 

Work,  net,  of  pump,  definition,  for- 
mula        24 

of  pump  in  horse-power 25 

Working  pressures,  wrought  iron  and 

steel  pipe,  data 367 

Worthington  Company,  forced  draft 
cooling   tower   with   surface 

condenser 354 

cooling  tower,  illustration 339 

mixing  chamber 352 

forced  draft,  cooling  tower 353 

hydraulic  tube-cleaner 324 

independent  jet  condenser,  illus- 
tration     290 

Worthington   Pump   and   Machinery 
Corporation,          centrifugal 

pump 110 

thrust  bearing. 121 

standard    condenser-tube    gland  302 
Wright      baffle-plate      exhaust-head, 

illustration 399 

Wrought  iron  and  steel  pipe,  working 

pressures,  data 367 

pipe,  trade  meaning 367 


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